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WORKS  OF  PROF.  J.  A.  MOVER 

PUBLISHED   BY 

JOHN  WILEY   &  SONS 


Descriptive  Geometry  for  Students  of  Engineering 

Third  Edition.  8vo,  viii  +  204  pages,  128  figures. 
Cloth,  $2.00. 

The  Steam  Turbine 

A  Practical  and  Theoretical  Treatise  for  Engineers 
and  Designers,  including  a  Discussion  of  the  Gas 
Turbine.  Second  Edition,  Revised  and  Enlarged. 
8vo,  ri  +  430  pages,  225  figures.  Gloth,  $4.00. 


STEAM  TURBINES 


A   PRACTICAL   AND    THEORETICAL 

TREATISE    FOR    ENGINEERS 

AND    DESIGNERS 

INCLUDING   A   DISCUSSION   OF  THE 
GAS   TURBINE 


BY  < 

JAMES  AMBROSE  MOVER,  S.B.,  A.M. 

MEMBER    OF    THE    AMERICAN    SOCIETY    OF    MECHANICAL    ENGINEERS,    MITGLIED    DBS 
VEREINES    DEUTSCHER   INGENIEURE,   MEMBRE  TITULAIRE  ASSOCIATION  INTER- 
NATIONALE    DU     FROID,     MEMBER      OF     THE     FRANKLIN     INSTITUTE, 
AMERICAN      INSTITUTE    OF     ELECTRICAL     ENGINEERS,     ETC. 

PROFESSOR  IN  CHARGE  OF  DEPARTMENT  OF  MECHANICAL  ENGINEERING,  THE, 
PENNSYLVANIA  STATE  COLLEGE,  FORMERLY  ENGINEER,  WESTING- 
HOUSE,  CHURCH,  KERR  &  CO.,  AND  ENGINEER  STEAM 
TURBINE  DEPARTMENT,   GENERAL 
ELECTRIC  COMPANY 


SECOND  EDITION,  REVISED  AND  ENLARGED 

FIRST  THOUSAND 


NEW  YORK 

JOHN    WILEY   &    SONS,   INC. 

LONDON:   CHAPMAN   &   HALL,   LIMITED 

1914 


COPYRIGHT,  1908, 1914 

"V 

JAMES  AMBROSE  MOVER 


Stanbopc  ipre&s 

F.    H.   G1LSON     COMPANY 
BOSTON.     U.S.A. 


Ml 


Library 


PREFACE    TO    FIRST    EDITION. 


THE  object  of  this  book  is  to  give  in  a  small  volume  what  I 
believe,  as  the  result  of  years  of  practical  experience,  engineers 
and  students  of  engineering  want  to  know  about  steam  turbines. 
It  is  intended  that  it  shall  be  a  manual  for  the  practical  engineer 
who  is  designing,  operating,  or  manufacturing  steam  turbines 
rather  than  a  compilation  of  manufacturers'  catalogs  combined 
with  a  digest  of  standard  books  on  thermodynamics  and 
mechanics. 

In  a  general  way  the  author  has  tried  to  explain  briefly  and 
directly  some  of  the  more  important  problems  about  which  the 
qualified  steam  engineer  must  have  some  knowledge.  When  this 
book  was  first  planned  it  was  intended  primarily  for  the  use  of  the 
author's  assistants  in  the  experimental  and  testing  departments 
of  one  of  the  large  manufacturing  companies,  but  later  it  seemed 
that  it  might  be  useful  in  a  larger  field. 

The  order  in  developing  the  subject  is  the  reverse  of  that  adopted 
by  most  authors.  Instead  of  discussing  the  intricacies  of  blading 
in  the  beginning  of  the  book,  the  more  simple  problems  of  nozzle 
design  are  presented  first.  A  great  deal  more  is  now  known  about 
nozzles  than  there  was  even  very  few  years  ago,  and  many  of  the 
conditions  affecting  the  efficiency  of  nozzles  may  now  be  considered 
well  established.  Nozzles  are  also  becoming  a  more  important 
part  of  all  types  of  turbines.  Even  the  Parsons  turbine  is  now 
being  modified  in  America  and  England  so  that  in  many  of  the 
latest  designs  for  large  sizes,  nozzles  are  used  in  the  high-pressure 
stages.  It  is  coming  to  be  generally  recognized  that  in  the  future 
there  will  probably  be  no  large  installations  of  reciprocating 
engines  for  electric  services.  A  few  years  ago  this  might  have 
been  considered  a  bold  statement,  but  it  is  a  fact  which  is  now 

iii 


291520 


iv  PREFACE 

generally,   although   reluctantly,  admitted   by   manufacturers    of 
reciprocating  engines. 

The  entropy-total  heat  chart  in  the  back  of  the  book  is  laid  out 
with  lines  of  constant  superheat  instead  of  lines  of  constant  tem- 
perature which  have  been  generally  used  for  charts  of  this  kind. 
For  practical  engineering  work  it  is  very  desirable  to  have  lines 
of  constant  superheat  on  such  charts,  because  in  America  and 
England  guarantees  of  steam  consumption  are  usually  given  in 
degrees  of  superheat  rather  than  of  temperature.  When  charts 
made  with  constant-temperature  lines  are  used,  it  is  always  neces- 
sary to  calculate  the  temperature  before  the  chart  can  be  used. 

Most  of  the  graduates  of  our  American  technical  schools  are 
entirely  "at  sea"  with  the  simplest  heat  calculations,  and  one  of 
the  reasons  for  this  deficiency  is  that  most  of  the  books  on  steam 
engines  —  and  especially  those  on  the  steam  turbine  —  are  more 
devoted  to  giving  a  large  quantity  of  facts  than  to  fulfilling  a  useful 
purpose.  Practical  engineers  who  have  had  to  deal  with  large 
numbers  of  men  with  an  engineering  training  agree  most  candidly 
with  Dr.  Steinmetz  when  he  says  in  substance  that  it  seems  to  cause 
no  concern  in  some  of  our  large  technical  schools  that  the  graduates 
are  sent  out  loaded  with  a  mass  of  half-understood  and  undigested 
subjects,  while  they  are  deficient  both  in  the  understanding  of  the 
fundamental  principles  and  in  the  ability  to  think.  If  this  volume 
can  serve  the  purpose  of  encouraging  students  to  think  it  will 
have  accomplished  one  of  its  principal  purposes,  not  losing  sight 
of  the  fact  that  the  book  is  intended  primarily  to  show  how  to 
do  things.4 

Nearly  all  the  proof-reading  has  been  done  by  Professor  John  F. 
Pelly  of  Philadelphia.  Because  of  Professor  Pelly's  thoroughly 
practical  as  well  as  theoretical  knowledge  of  the  subject  matter,  his 
conscientious  and  painstaking  work  is  very  greatly  appreciated. 

I  take  this  opportunity  to  thank  Professor  Ira  N.  Hollis  and 
Professor  F.  Lowell  Kennedy  of  Cambridge  for  the  criticisms  and 
suggestions  which  I  received  from  them  when  the  manuscript  of 
this  book  was  preparing.  I  am  also  greatly  indebted  to  Mr.  Walter 
C.  Kerr,  president,  and  Mr.  Sidney  E.  Junkins,  vice-president  of 


PREFACE  V 

Westinghouse,  Church,  Kerr  &  Company,  for  their  encouragement 
and  for  making  it  possible  to  finish  the  book  at  this  time. 

For  placing  at  my  disposal  a  great  deal  of  information  regarding 
the  latest  results  in  steam  turbine  engineering,  which  is  usually 
very  difficult  to  obtain,  I  am  particularly  indebted  to  Mr.  Richard 
H.  Rice  of  Lynn,  and  Mr.  J.  R.  Bibbins  of  Pittsburg. 

I  wish  to  thank  Professor  Arthur  M.  Greene  of  Troy  and 
Mr.  Albert  Stritmatter  of  Cincinnati  for  suggestions  relating  to  the 
subject  matter.  For  various  services  in  the  preparation  of  this 
book,  I  should  mention  also  Messrs.  Francis  Hodgkinson  and 
Harold  P.  Childs  of  the  Westinghouse  Machine  Company;  C.  P. 
Crissey,  S.  A.  Moss,  and  W.  E.  Culbertson  of  the  General  Electric 
Company;  C.  P.  Chasteney  of  the  De  Laval  Steam  Turbine 
Company;  James  Wilkinson,  president  of  the  Wilkinson  Turbine 
Company;  St.  John  Chilton  of  the  Allis-Chalmers  Company;  H.  H. 
Wait  of  the  Western  Electric  Company;  Carl  S.  Dow  of  the  B.  F. 
Sturtevant  Company;  and  J.  Clarence  Moyer  of  Philadelphia. 

Many  of  the  illustrations  for  the  book  have  been  provided,  in 
some  cases  at  considerable  expense  to  themselves,  by  the  Gassier 
Magazine  Company,  Westinghouse  Machine  Company,  General 
Electric  Company,  De  Laval  Steam  Turbine  Company,  Rateau 
Turbine  Company,  Kerr  Turbine  Company,  Wilkinson  Turbine 
Company,  Allis-Chalmers  Company,  C.  H.  Parsons  &  Co.,  and 
Brown,  Boveri  &  Co. 

Throughout  the  text  important  words  and  sentences  are  brought 
out  by  the  use  of  bold-faced  type,  thus  making  the  subjects  of  a 
paragraph  visible  at  a  glance. 

The  author  is  always  glad  to  answer  correspondence  with 
teachers  relating  to  questions  which  inevitably  arise  in  the  discussion 
of  designs  for  steam  turbines,  all  of  which  cannot,  of  course,  be 
taken  up  in  detail  in  any  book. 

JAMES  AMBROSE  MOYER. 

417  WEST  n8xH  ST.,  NEW  YORK, 
September,   1908. 


PREFACE  TO    SECOND    EDITION. 


SINCE  the  issuance  of  the  first  edition,  scientific  investigation 
has  not  added  much  to  our  previous  knowledge  of  the  proper- 
ties of  steam,  nor  have  the  new  types  invented  in  the  interval 
become  commercially  successful.  In  fact  the  trend  of  things 
has  been  rather  toward  the  general  adoption  of  one  of  four 
types  of  steam  turbines:  (i)  a  single  turbine  wheel  of  the  im- 
pulse type;  (2)  impulse  wheels  with  two  velocity  stages  in 
each  pressure  stage;  (3)  drum  construction  with  "  reaction  " 
blading;  and  (4)  a  combination  of  (2)  and  (3)  called  a  combined 
impulse  and  reaction  type.  Discoveries  like  that  of  Tesla's,  claim- 
ing to  have  made  possible  very  great  simplification  of  turbine 
construction  with  unheard  of  improvements  in  economy,  have 
at  times  attracted  the  attention  of  engineers,  but  always  with 
the  final  result  that  the  claims  have  not  been  made  good. 

The  really  important  developments  of  the  last  few  years  have 
been  in  the  construction  of  increasingly  large  sizes.  The  largest 
turbine-generator  now  ready  for  installation  is  rated  at  35,000 
kilowatts,  which  is  to  be  compared  with  a  maximum  size  of 
14,000  kilowatts  of  only  three  years  ago.  Certainly  these  are 
the  days  typical  of  the  concentration  of  power  in  large  units, 
not  only  in  the  turbine  and  generator  room,  but  also  in  the  boiler 
room,  in  the  condenser  pit,  and  in  the  installation  of  the  other 
plant  auxiliaries.  Extremely  large  sizes  are  being  installed  be- 
cause they  effect  a  substantial  reduction  in  the  unit  cost  of 
power  generation.  In  spite  of  the  general  increase  in  the  cost 
of  raw  and  manufactured  materials  in  the  last  five  years,  the 
application  of  steam  turbines  in  power  plants  in  the  place  of 
reciprocating  engines  has  reduced  the  total  first  cost  of  large 
first-class  power  plants  from  $120  per  kilowatt  of  rated  capacity, 

vii 


viii  PREFACE  TO  SECOND   EDITION 

which  was  a  fair  average  value  five  years'  ago,  to  nearly  $60 
to-day. 

In  these  times  when  there  is  such  general  discussion  of  con- 
servation and  efficiency,  the  low-pressure  steam  turbine  takes 
an  important  place,  because  of  its  innumerable  applications  for 
preventing  the  wasting  of  any  steam  to  the  atmosphere.  In 
nearly  every  large  central  station  hundreds  of  pounds  of  steam 
exhausted  from  the  auxiliaries  at  atmospheric  pressure  in  excess 
of  that  required  in  the  heaters  for  heating  the  feed  is  lost  through 
the  exhaust-heads.  Modern  methods  of  turbine  application 
would  save  and  utilize  this  steam  for  power.  In  the  most 
modern  practice,  therefore,  the  greatest  skill  of  the  engineer  is 
called  upon  more  in  connection  with  the  methods  of  applying 
the  commercial  types  of  turbines  already  developed  rather  than 
in  the  actual  designing  of  new  types  of  machines.  In  steam 
engine  designing  there  have  been  always  unlimited  possibilities; 
in  steam  turbine  designing  these  are  few. 

Most  of  the  additions  made  in  this  edition  have  been,  there- 
fore, mainly  in  the  line  of  new  applications.  The  chapter  on 
low-pressure  turbines  has  been  rewritten  and  very  much  extended 
to  include  the  latest  developments  and  applications.  This 
chapter  should  be  unusually  interesting  to  all  engineers  and 
students.  New  chapters  have  been  added  on  Bleeder  or  Ex- 
traction Turbines  and  Mixed  Pressure  Turbines.  Both  mark 
recent  successful  developments  in  turbine  applications,  making 
it  a  still  more  important  competitor  of  the  reciprocating  engine 
in  the  non-condensing  field.  The  chapters  on  Heat  Theory, 
Steam  Flow,  Nozzle  Design,  Blade  Design  and  Reaction  Tur- 
bine Design  have  been  rewritten  with  the  addition  throughout 
the  text  of  many  illustrative  examples  and  the  inclusion  in  an 
appendix  of  a  large  number  of  practical  exercises  and  problems 
to  illustrate  important  principles,  thus  making  the  book  con- 
siderably more  serviceable  than  before  as  a  class-room  text. 
For  these  exercises  the  data  are  selected  in  most  cases  so  as  to 
simplify  the  calculations  and  to  avoid  taking  too  "much  of  the 
tune  of  the  student  or  reader  with  purely  numerical  work.  A 


PREFACE   TO   SECOND   EDITION  ix 

new  entropy-total  heat  chart  has  been  calculated  and  engraved, 
which  embodies  the  most  recent  and  reliable  data  on  the  prop- 
erties of  superheated  and  saturated  steam. 

In  the  preparation  of  this  edition  I  am  particularly  indebted 
to  Professor  J.  E.  Enswiler  of  the  University  of  Michigan  and 
Professor  J.  P.  Calderwood  of  The  Pennsylvania  State  College 
for  innumerable  suggestions. 

Many  important  suggestions  and  criticisms  have  been  re- 
ceived from  Professor  J.  V.  Ludy,  Purdue  University;  Professor 
E.  A.  Fessenden,  University  of  Missouri;  Mr.  M.  Nusim,  Gen- 
eral Electric  Company,  Lynn,  Mass.;  Mr.  C.  P.  Crissly,  Henry 
Worthington  Co.,  New  York;  Mr.  C.  P.  Chasteney,  De  Laval 
Steam  Turbine  Co.;  and  Mr.  N.  C.  Miller  of  The  Pennsylvania 
State  College. 

Finally,  important  mention  must  be  made  of  assistance  in  this 
work  received  from  Mr.  H.  T.  Herr,  Vice-President  and  General 
Manager  of  the  Westinghouse  Machine  Co.;  Mr.  Francis  Hodg- 
kinson  of  the  same  company;  Messrs.  Richard  H.  Rice  and 
Chas.  K.  West  of  the  General  Electric  Company;  Dr.  E.  J. 
Berg  of  Union  College;  Mr.  Alfred  Rigling  of  The  Franklin 
Institute;  and  The  Electric  Journal. 

THE  AUTHOR. 

STATE  COLLEGE,  PA., 
January  i,  1914. 


CONTENTS 


CHAPTER  PAGE 

I.     INTRODUCTION i 

II.     THE  ELEMENTARY  THEORY  OF  HEAT 10 

III.  NOZZLE  DESIGN 29 

IV.  STEAM  TURBINE  TYPES  AND  BLADE  DESIGN 58 

V.     MECHANICAL  LOSSES  IN  TURBINES 115 

VI.     METHOD  FOR  CORRECTING  STEAM  TURBINE  TESTS 126 

VII.     COMMERCIAL  TYPES  OF  TURBINES 140 

De  Laval,  Parsons,  Westinghouse,  Allis-Chalmers,  Curtis,  Rateau, 
Wilkinson,  Zoelly,  Sturtevant,  Riedler-Stumpf,  Kerr,  Terry, 
Dake,  etc. 

VIII.     GOVERNING  STEAM  TURBINES 218 

IX.     LOW-PRESSURE  (EXHAUST)  TURBINES 247 

X.    MIXED-PRESSURE  TURBINES 2$8a 

XI.    BLEEDER  OR  EXTRACTION  TURBINES 2586 

XII.    MARINE  TURBINES 259 

XIII.  TESTS  OF  TURBINES 261 

XIV.  STEAM  TURBINE  ECONOMICS 273 

XV.    STRESSES  IN  RINGS,  DRUMS,  AND  DISKS 313 

Design  of  Turbine  Wheels  —  Critical  Speeds  of  Loaded  Shafts. 

XVI.    GAS  TURBINES 340 

XVII.    ELECTRIC  GENERATORS  FOR  TURBINES 358 

Direct-current  Generators  —  Alternating-current  Generators. 

APPENDIX  OF  PRACTICAL  EXERCISES 363 


THE   STEAM   TURBINE 


CHAPTER    I. 
INTRODUCTION. 

THE  steam  turbine  is  the  most  modern  as  well  as  the  most 
ancient  steam  motor.  Recently  its  development  has  gone  by 
leaps  and  bounds;  and,  above  all,  in  its  applications  it  is  gaining 
ground  daily.  Doubtless  it  is  to  be  the  most  important  prime 
mover  of  the  near  future. 

During  recent  years  results  have  been  secured  with  steam  tur- 
bines that  only  a  short  time  ago  were  considered  practically  unat- 
tainable. Primarily  their  great  success  lies  in  their  adaptability 
to  operation  with  high  vacuums.  Steam  turbines  are,  therefore, 
almost  ideally  suitable  for  the  conditions  of  modern  engineering 
practice  requiring  both  high  vacuums*  and  high  superheats. 
To-day  in  the  economical  use  of  steam  they  are  unrivaled;  and, 
because  of  improved  manufacturing  methods,  marking  the  tran- 
sition from  the  experimental  to  the  commercial  stage,  first  cost 
is  no  longer  a  deciding  factor  favoring  reciprocating  engines. 

Compared  with  reciprocating  steam  and  gas  engines,  steam 
turbines  require  much  smaller  and  cheaper  foundations,  occupy 
less  floor  space,  require  fewer  attendants,  and  because  no  lubrica- 
tion is  required  for  any  parts  in  contact  with  the  steam,  the  con- 
densation becomes  directly  available  for  feed  water.  The  highest 
superheats  can  be  employed  without  affecting  the  choice  of  lubri- 
cants, and  the  cost  of  oil  for  lubrication  is  very  low. 

A  steam  turbine  of  the  simplest  type  is  essentially  a  wheel 
similar  to  an  ordinary  water  wheel,  which  is  moved  around  by 
a  steam  jet  impinging  on  its  blades.  Steam  is  directed  against 
the  turbine  wheel  by  nozzles  or  similar  passages  delivering  the 

*  The  question  of  the  most  profitable  vacuum  for  given  conditions  is  discussed 
on  pages  273  to  281. 

I 


THE   STEAM    TURBINE 


FIG.   i.  —  A  Small  Modern  Steam  Turbine  with  Part  of  the  Casting  Removed. 

steam  at  mathematically  exact  angles,  calculated   to  make  the 
steam  strike  the  blades  of  the  wheel  most  advantageously. 

Fig.  i  is  an  illustration  of  a  modern  steam  turbine  with  a  part 
of  the  casing  removed  to  show  the  construction.     The  turbine 


INTRODUCTION 


wheel  W  is  shown  here  with  numerous  blades  on  its  circumference. 
The  steam  comes  to  the  turbine  frorn  the  boilers  through  a  suitable 
steam  main  connected  to  the  top  of  the  turbine  at  M  and  passes 
down  through  the  pipe  A  to  the  steam-chest  B.  From  this  steam- 


FIG.  2.— The  Turbine  Wheel  and  Nozzles. 

chest  it  is  guided  through  one  or  more  nozzles,  from  which  it 
escapes  at  a  high  velocity  to  impinge  on  the  blades  on  the  circum- 
ference of  the  turbine  wheel,  which  is  thus  made  to  rotate,  and 
performs  work  by  moving  machinery  connected  to  the  shaft. 
Nozzles  frorn  which  the  steam  is  discharged  are  located  around 
the  periphery  of  the  wheel  as  shown  in  Fig.  2  with  their  enlarged 


4 


THE    STEAM   TURBINE 


M 


ends,  technically  called  mouths,  very  close  to  the  blades.*  Steam 
after  passing  through  the  blades  enters  the  exhaust  pipe  at  E 
(Fig.  i )  and  is  discharged  into  the  atmosphere  or  into  a  condenser, 
depending  on  whether  the  operation  is  non-condensing  or  con- 
densing. 

Preliminary  to  the  study  of  the  modern  commercial  types  of 
steam  turbines  it  is  desirable  to  state  briefly  some  of  the  most 
important  stages  through  which  this  very  ancient  form  of  steam 
motor  has  passed  in  its  development. 

Early  History.  The  earliest  notices  of  heat  engines  of  any 
kind  are  found  in  a  book  by  Hero  of  Alexandria,  which  was 
probably  written  in  the  second  century  before  Christ.  In  this 
book  of  mechanical  contrivances  a  steam  reaction  wheel  is  men- 
tioned. This  first  steam  turbine  is  shown 
in  Fig.  3.  It  is  described  as  consisting 
of  a  hollow  spherical  vessel  pivoted 
on  a  central  axis  and  supplied  with 
steam  through  the  support  M  and  one 
of  the  pivots  from  a  boiler,  B,  beneath. 
Steam  escaped  from  the  vessel  through 
bent  pipes  or  nozzles  N,  N,  facing  tan- 
gentially  in  opposite  directions.  The 
spherical  vessel  was  revolved  by  the  re- 
action due  to  the  escaping  steam,  just  as 
a  "  Barker's  mill  "  is  moved  by  the  water 
escaping  from  its  arms.  Any  fluid  escap- 
ing under  pressure  from  a  vessel  which 
is  free  to  move  causes  a  "reaction"  tending  to  displace  the  vessel 
in  the  opposite  direction  from  the  flow  of  the  fluid.  This  reaction, 
although  imperfectly  understood  by  Hero,  was  perfectly  applied 
in  his  steam  turbine  which  was  used  to  open  the  doors  of 
temples.  Only  a  few  years  ago  a  model  of  Hero's  engine  was 
constructed  by  a  celebrated  English  engineer,!  with,  of  course, 

*  In  this  figure  one  of  the  nozzles  is  represented  as  if  transparent  to  show  its 
shape  on  the  inside,  and  a  part  of  the  steel  band  around  the  blades  is  cut  away  to 
show  the  shape  of  the  blades  or  vanes,  as  well  as  to  illustrate  the  passage  of  steam 
from  the  nozzle  into  and  through  the  blades. 

t  See  page  8. 


FIG.  3.  —  Hero's  Turbine. 


INTRODUCTION  5 

all  the  advantages  of  modern  machine  tools  and  appliances, 
with  the  result  that  an  engine  was  produced  which,  in  economy, 
compared  well  with  our  elaborate  and  complicated  modern 
engines. 

In  1577  a  German  mechanician,  it  is  said,  used  a  turbine  simi- 
lar to  Hero's  to  rotate  reaming  and  burnishing  tools,  but  from 
the  time  of  Hero  down  to  the  seventeenth  century  there  is  no 
record  of  progress  in  the  development  of  steam  heat  engines.  In 
1629  Branca,  an  Italian  architect,  designed  a  steam  turbine 
(Fig.  4)  resembling  a  water  wheel,  which  was  driven  by  the 
impulse  from  a  jet  of  steam  directed  by  means  of  a  nozzle  upon 
suitable  vanes  attached  to  the 
wheel.  Branca's  turbine  en- 
gine, however,  was  not  success- 
ful; and  until  the  end  of  the 
nineteenth  century,  although  in 
the  interval  many  steam  tur- 
bines and  other  rotary  engines 
were  patented,  the  piston  or 
reciprocating  steam  engine,  FIG.  4.  —  Branca's  Turbine, 

under  the  leadership  of  Watt, 

had,  commercially,  an  unrestricted  field  and  remarkable  results 
were  accomplished. 

It  is  interesting  to  observe  that  the  modern  type  of  impulse 
turbine  with  a  single  row  of  blades  like  the  one  illustrated  in 
Fig.  i  is  practically  the  same,  except  for  details,  as  the  historic 
Branca's  wheel.  The  principal  difference  is  that  Branca's  wheel 
was  not  enclosed  in  a  casing.  Essential  parts  —  the  nozzle, 
the  blades,  the  wheel,  and  the  shaft  —  were  practically  the  same 
as  in  some  modern  machines.  Probably  if  Branca  had  under- 
stood the  laws  of  the  expansion  of  steam  as  we  do  to-day,  he  could 
have  made  a  successful  prime  mover  of  his  turbine.  Those  who 
came  after  him  were  aided  not  only  by  a  superior  knowledge 
but  also  by  the  opportunities  for  scientific  investigation  and  the 
skill  of  our  present-day  workshops. 

De  Laval  Type.     Dr.  Gustaf  De  Laval,  a  Swedish  scientist,  was 


6  THE   STEAM   TURBINE 

a  pioneer  in  the  modern  commercial  development  of  steam  tur- 
bines. In  1882  he  constructed  his  first  steam  turbine,  which  was 
similar  in  principle  to  Hero's  reaction  engine.  De  Laval's  first 
turbine  was  designed  primarily  for  driving  his  milk  and  cream 
separators,  for  which  there  was  then  a  large  sale.  For  other 
purposes,  however,  there  was  no  general  application,  because  at 
the  very  high  speeds  for  which  they  were  designed,  it  was  difficult 
to  utilize  the  power;  and  besides,  the  steam  consumption  was 
practically  prohibitive. 

Later  De  Laval  turned  his  attention  to  the  development  of 
Branca's  steam  turbine,  and  was  remarkably  successful.  After 
much  experimenting,  he  developed  an  impulse  turbine  which  is 
still  one  of  the  standard  makes.  (See  Figs.  82  to  86.)  This 
great  engineer,  after  investigating  the  possibilities  of  both  Hero's 
and  Branca's  types  and  having  decided  to  adopt  the  latter,  began 
then  some  strikingly  original  inventive  work,  which,  in  many 
respects,  led  the  way  for  the  accomplishments  of  to-day. 

It  should  be  stated,  however,  in  this  connection,  that  no  engineer 
thinks  of  belittling  De  Laval's  work  because  his  investigations 
were  mostly  in  the  line  of  improvements  to  existing  types.  Un- 
questionably he  must  have  the  credit  for  producing  the  first 
commercially  successful  steam  turbine.  Many  of  the  features  of 
his  original  designs  have  actually  contributed  in  no  small  meas- 
ure to  our  knowledge  of  machine  design  and  thermodynamics, 
and  have  become  fundamental  principles  underlying  many  of  the 
most  important  modern  steam  turbine  developments.* 

Parsons  Type.  With  the  early  work  of  De  Laval,  however, 
the  development  of  steam  turbines  designed  to  operate  by  the 
reaction  principle  of  Hero's  engine  was  not  given  up.  Almost 
contemporaneously  with  De  Laval,  C.  A.  Parsons  in  England 
began  the  development  of  the  well-known  type  which  to-day  bears 
his  name,  and  which  has  made  possible  the  brilliant  records  of 
turbine  ocean  steamers.  In  April,  1884,  this  great  inventor  took 

*  The  most  important  feature  introduced  by  De  Laval  is  that  of  the  diverging 
nozzle  (British  Patent  No.  7143  of  1889),  the  principle  of  which  has  influenced 
the  development  of  practically  all  types  of  steam  turbines. 


INTRODUCTION  / 

out  his  first  patents  on  steam  turbines.  The  practicability  of  the 
steam  turbine  he  then  proposed  is  a  striking  feature  of  even  his 
first  patents.  His  specifications  showed,  above  all,  that  a  great 
deal  of  time  and  thought  had  been  devoted  to  constructive  details. 
Methods  for  reducing  vibration,  preventing  leakage  of  steam,  and 
providing  for  efficient  lubrication  contributed  very  largely  to  his 
success.  Many  of  the  details  of  this  early  turbine  are  now  obso- 
lete, so  that  only  a  very  short  description  will  be  given  here.  A 
section  drawing  of  Parsons*  first  turbine  is  shown  in  Fig.  5.  A 


FIG.  5.  — Early  Parsons  Steam  Turbine. 

large  central  collar,  C,  is  attached  to  the  main  shaft,  S,  which  runs 
the  length  of  the  turbine.  At  the  ends  of  the  casing  where  the 
shaft  passes  through  it  the  cross-section  is  reduced.  The  main 
shaft,  S,  supports  a  large  number  of  rings  which  are  held  in  place 
between  the  collar,  C,  and  the  nuts,  N,  which  are  screwed  on  the 
reduced  section  of  the  shaft  at  the  ends.  These  rings,  around 
their  circumferences,  support  those  turbine  blades  (b,  b)  which 
move  with  the  shaft.  There  are,  however,  alternating  with  them, 
other  rows  of  blades  (c,  c)  attached  to  the  inside  of  the  turbine 
casing.  Technically  the  blades  b,  b  are  called  moving  blades, 
and  c,  c  are  called  fixed  or  stationary  blades.  Steam  is  admitted 
to  the  turbine  blades  through  the  annular  chamber,  A,  encircling 
the  collar,  C,  and  then  it  passes  to  the  right  and  left  through  the 
alternate  rows  of  stationary  and  moving  blades  to  the  exhaust 
passages  E,  E  —  one  at  each  end  of  the  turbine.  The  steam 
expands  in  the  blades  as  in  a  nozzle,  and  its  reaction  moves  the 
blades  attached  to  the  shaft,  just  as  Hero's  turbine  was  rotated 
by  the  steam  escaping  from  its  arms. 


8 


THE   STEAM   TURBINE 


By  the  "double-flow"  arrangement  in  this  design  by  which  the 
steam  is  passed  from  the  center  to  the  exhaust  at  both  ends  there 
can  be  very  little  axial  thrust  on  the  shaft.  Any  thrust  that 
does  occur,  however,  is  balanced  by  the  pressure  of  the  exhaust 
steam  in  the  chambers  E,  E  at  the  ends  of  the  casing.  A  slight 
movement  of  the  shaft  toward  either  end  checks  the  flow  of  the 
exhaust  steam,  and  increases  the  back  pressure  at  that  end.  This 
increased  pressure  then  moves  the  shaft  back  to  its  normal 
position. 

Usually  it  is  not  possible  to  balance  the  parts  of  a  rotating  mass 
to  make  its  center  of  gravity  coincide  exactly  with  the  geometric 
center  about  which  it  revolves.  In  any  machine  like  a  steam 
turbine,  when  these  two  centers  do  not  coincide  excessive  vibra- 
tions of  the  shaft  are  produced  which  at  certain  speeds  *  are 
sufficient  to  break  it.  To  overcome  this  difficulty,  Parsons  in- 
geniously allowed  a  little  lateral  play,  or  "  elasticity,"  as  he  called 


i — i 


FIG.  6. —  Screw  Type  of  Steam  Turbine. 

it,  for  the  shaft  by  means  of  a  series  of  rings  of  two  different 
diameters,  in  principle  very  much  the  same  as  the  present  con- 
struction of  the  main  bearings  of  Parsons  turbines  (see  Fig.  100), 
so  that  it  was  permitted  to  move  laterally  a  certain  amount,  say  a 
hundredth  of  an  inch,  to  allow  the  proper  adjustment  in  passing 
from  rest  to  the  normal  speed  of  running. 

Among  his  early  experiments  Parsons  also  tried  a  purely 
reaction  steam  turbine,  following  almost  exactly  the  published 
designs  of  Hero.  This  turbine,  running  with  100  pounds  per 

*  This  phenomenon  occurs  at  very  definite  speeds,  called  "critical,"  for  every 
rotating  mass.  Fuller  discussion,  with  a  method  for  calculating  "critical"  speeds, 
is  given  on  page  338. 


INTRODUCTION  9 

square  inch  steam  pressure  and  27  inches  of  vacuum,  gave  an 
output  of  20  horsepower  at  5,000  revolutions  per  minute.  Steam 
consumption  was  only  40  pounds  per  brake  horsepower  per 
hour,  which  was  indeed  a  remarkably  good  result  for  that  time. 

Screw  Type.  Still  another  kind  of  turbine,  of  only  historical 
interest,  should  be  mentioned.  A  large  number  of  inventors  have 
worked  on  the  development  of  a  screw  type  like  Fig.  6.  Hewitt 
worked  for  a  long  time  on  a  turbine  of  this  kind,  and  finally  con- 
cluded the  results  were  not  satisfactory.  Steam  was  admitted  to 
this  turbine  through  the  chamber  A,  and  passed  through  holes  in 
the  plates  P,  P  into  the  helical  grooves  on  the  shaft.  In  these 
grooves  the  steam  was  expanded  and  then  escaped  to  the  exhaust 
pipes  E,  E  at  the  two  ends.  Effective  action  of  the  steam  was 
probably  obtained  only  in  the  first  part  of  the  grooves;  and  after 
being  deflected  into  a  helical  course,  it  rushed  through  to  the 
exhaust  without  much  additional  effect  in  moving  the  shaft. 
Excessive  leakage  of  steam  between  the  helical  threads  and  the 
casing  was  another  serious  difficulty. 

Recently  a  somewhat  similar  arrangement  having  two  "  screw 
wheels  "  meshing  together  not  unlike  spiral  or  helical  gears  has 
been  successfully  developed  by  the  Buffalo  Forge  Company 
(see  Figs.  157  and  158,  page  2i7a). 


CHAPTER   II. 
THE  ELEMENTARY  THEORY  OF  HEAT. 

NOTE.  —  This  short  chapter  may  well  be  omitted,  in  reading,  by  those  who  are 
familiar  with  the  thermodynamics  of  heat  engines  and  with  the  use  of  entropy 
diagrams.  It  is  intended  primarily  for  practical  engineers,  who  will  find  it  par- 
ticularly valuable  for  reference "  purposes,  as  the  subject  matter  is  completely 
indexed. 

TECHNICALLY  the  steam  turbine  must  be  regarded  as  a  heat 
engine,  that  is,  a  machine  in  which  heat  is  employed  to  do  mechan- 
ical work.  From  the  viewpoint  of  the  practical  man  its  function, 
the  same  as  that  of  any  other  heat  motor,  is  to  secure  as  much 
work  as  possible  from  a  given  amount  of  steam,  or,  going  a  step 
farther  back,  from  the  combustion  of  a  given  amount  of  fuel. 
Heat  theory  is,  therefore,  of  first  importance. 

Heat  is  a  form  of  energy  like  electrical,  chemical,  mechanical, 
potential,  and  kinetic.  No  doubt  exists  about  the  equivalence 
of  the  different  forms  of  energy  and  their  close  relation  to  each 
other.  Each,  at  will,  can  be  changed  into  any  of  the  other  forms. 

The  relative  amount  of  heat  in  a  body  is  observed,  in  common 
experience,  by  the  sense  of  touch  —  whether  the  body  is  a  solid, 
a  liquid,  or  a  gas.  By  such  experience  we  have  learned  to  recog- 
nize certain  sensations  as  hot  or  cold;  and  then,  with  more 
accuracy,  to  speak  of  degrees  of  temperature.  Now  when  a  hot 
and  a  cold  body  are  brought  together  their  temperatures  become 
equalized.  The  hotter  body  always  loses  heat.  The  colder 
body  always  gains  heat.*  This  experience  is  the  principal  basis 
for  all  heat  calculations. 

When  in  the  course  of  time  it  had  been  found  that  a  more 
accurate  method  than  that  of  the  sense  of  touch  was  needed  for 
heat  determinations,  methods  utilizing  the  expansion  of  liquids 

*  This  phenomenon  is  called  the  second  law  of  thermodynamics,  —  that  "heat 
energy  always  passes  from  a  warm  body  to  a  cold  body. " 

10 


THE  ELEMENTARY  THEORY  OF  HEAT       II 

came  to  be  generally  employed.  Many  substances  have  a  practi- 
cally uniform  rate  of  expansion  between  the  limits  of  temperature 
an  engineer  has  to  deal  with.  A  small  column  of  mercury  in  a 
glass  capillary  tube  is  usually  taken  as  a  standard  for  temperature 
measurements.*  The  mercury  in  an  accurate  thermometer 
expands  very  nearly  ^||  of  its  volume  when  heated  from  the 
freezing  temperature  of  water  (32°F.)  to  the  boiling  point 
(2i2°F.).  The  expansion  between  the  freezing  point  and  the 
boiling  point  of  water  has  therefore  been  called,  arbitrarily, 
180°  F. 

For  theoretical  heat  calculations  the  zero  of  temperature  is 
taken  as  492°  F.  below  the  freezing  temperature  of  water;  or, 
460°  below  the  Fahrenheit  zero.  This  very  low  temperature  is 
called  the  absolute  zero,  and  at  this  point  there  is  theoretically 
no  heat  energy. 

Temperatures  measured  from  the  absolute  zero  are  called 
absolute  temperatures  and  are  indicated  generally  by  T,  to  dis- 
tinguish them  from  the  ordinary  Fahrenheit  temperatures,  t,  as 
read  on  a  thermometer  scale. 

Using  these  symbols,  we  have  then  in  Fahrenheit  degrees, 

T  =  t  +  460. 

Absolute  temperatures  are  convenient  for  heat  calculations 
because  "  perfect  "  gases,  at  constant  pressure,  increase  in  volume 
in  proportion  to  the  increase  in  absolute  temperature. 

*  The  ordinary  mercury  thermometers  can  be  used  to  measure  temperatures  to 
about  575°  F.  with  accuracy.  For  higher  temperatures  the  capillary  tube  over 
the  mercury  should  be  filled  with  nitrogen  or  carbonic  acid  gas  under  high  pressure. 
Such  thermometers  can  then  be  used  for  temperatures  up  to  1000°  F. 

If  the  mercury  is  not  throughout  its  whole  length  at  the  same  temperature  as 
that  being  measured,  a  correction,  k,  given  by  the  following  formula  must  be 
added  to  the  observed  temperature,  /,  in  Fahrenheit  degrees: 

k  =  .000,088  D(t  —  t'), 

where  D  is  the  length  of  the  mercury  column  exposed,  measured  in  Fahrenheit 
degrees,  and  t'  is  the  temperature  of  the  exposed  part  of  the  thermometer.  When 
long  thermometers  are  used  in  shallow  wells  in  high-pressure  steam  pipes  this  cor- 
rection is  often  5°  to  10°  F.  For  experimental  data  and  direct-reading  correction 
curves,  see  Moyer's  Power  Plant  Testing,  2d  edition  (McGraw-Hill  Book  Co.), 
pages  31-33. 


12  THE   STEAM    TURBINE 

Heat  Units  and  Specific  Heat.  The  amount  of  heat  required 
to  raise  the  temperature  of  one  pound  of  water  from  62°  to  63°  F. 
is  taken  arbitrarily  as  the  standard  English  unit  of  heat,  —  com- 
monly called  the  British  thermal  unit  (B.T.U.).*  The  ratio  of 
the  amount  of  heat  required  to  raise  the  temperature  of  a  pound 
of  water  or  steam  one  degree  to  the  British  thermal  unit  is  called 
the  specific  heat.f 

The  specific  heat  of  steam  and  of  gases  changes  in  value  accord- 
ing to  the  conditions  under  which  the  heat  is  applied.  If  heat  is 
added  to  a  vapor  or  a  gas  held  in  a  closed  vessel,  with  no  chance 
for  expansion,  no  external  work  is  done,  and  therefore  practically 
all  the  heat  added  is  used  to  increase  the  temperature.  This  is 
the  condition  in  a  boiler  when  no  steam  is  being  drawn  off.  In 
this  case  the  specific  heat  is  symbolized  by  Cv  =  specific  heat  at 
constant  volume.  If,  on  the  other  hand,  the  pressure  is  kept 
constant  but  the  volume  is  allowed  to  change  to  permit  expansion 
and  the  performing  of  external  work,  we  say  then,  Cp  =  specific 
heat  at  constant  pressure. 

Heating  at  constant  pressure  is  the  condition  that  is  most 
interesting  to  the  engineer.  When  his  engines  are  running  the 
boilers  are  making  steam  at  constant  pressure.  The  heat  energy 
absorbed  by  a  pound  of  steam  for  raising  only  the  temperature 
must  be,  obviously,  approximately  the  same,  regardless  of  the 
conditions  of  pressure  and  volume.  Since  for  constant  pressure 
conditions  some  external  work  is  always  done,  requiring  a 
larger  amount  of  heat  energy  than  for  the  case  when  the  volume 
is  constant,  it  follows  that  Cp  is  always  greater  than  Cv. 

We  should  add,  further,  that  an  engineer's  calculations  con- 
cerning energy  transformations  in  steam  turbines  are  almost 

*  In  the  C.  G.  S.  system  of  units  the  kilogram-calorie,  called  in  German 
Warmeeinheit  (WE),  is  used  as  the  standard  heat  unit,  i  kg.-cal.  or  i  WE  =  3.97 
or  nearly  4  B.T.U. 

t  The  specific  heat  of  water  at  200°  F.  is  1.005,  and  of  superheated  steam  an 
average  value  of  .6  is  often  assumed  in  rough  calculations  for  steam  at  the  usual  boiler 
pressures  in  power  plant  practice  for  superheats  less  than  150°  F.  Mean  values  of 
the  specific  heat  of  superheated  steam  are  given  by  the  curves  in  Fig.  30. 


THE  ELEMENTARY  THEORY  OF  HEAT       13 

without  exception  for  the  condition  of  constant  pressure,  and, 
consequently,  only  values  of  Cp  are  generally  useful.  Most  gases 
have  practically  constant  values  for  their  specific  heats. 

At  temperatures  near  the  boiling  point,  the  heating  of  vapors, 
like  steam,  is  influenced  by  molecular  attraction,  so  that  their 
specific  heats  are  variables  depending  on  conditions  of  tempera- 
ture and  pressure.  .  The  specific  heat  of  superheated  steam  de- 
creases with  increasing  temperatures  to  a  minimum  value.  The 
values  of  specific  heat  increase  slightly,  on  the  other  hand,  with  an 
increase  of  pressure.* 

Mechanical  Equivalent  of  Heat.  Heat  and  work  are  both 
forms  of  energy  and  are  "  equivalent,"  meaning  that  energy  can 
be  transformed  into  mechanical  work,  and  that  work,  as  a  form 
of  energy,  can  be  changed  back  again  into  heat.  The  relation  is 
expressed  by 

i  British  thermal  (heat)  unit  =  778  foot-pounds  (work). 

HEAT  AND  WORK. 

Heat  is  a  form  of  energy.  Each  of  the  various  kinds  of  heat 
motors,  such  as  the  steam  engine,  the  steam  turbine,  the  gas 
engine,  or  the  gas  turbine,  is  a  machine  for  obtaining  mechanical 
work  from  heat  energy. 

In  the  general  principles  of  operation  the  steam  turbine  and  the 
reciprocating  or  "  piston  "  steam  engine  are  essentially  similar 
machines.  Both  do  work  according  to  the  same  heat  relations. 
The  gas  turbine  is  somewhat  different.  This  new  motor,  which 
as  yet  has  scarcely  reached  a  practical  stage  of  development, 
will  be  discussed  in  its  proper  place. 

In  a  reciprocating  steam  engine  working  "  expansively  "  the 
steam  is  admitted  at  boiler  pressure  until  the  point  of  cut-off;  and 
during  the  remainder  of  the  stroke  the  piston  is  pushed  ahead,  or 
does  work,  by  the  expansion  of  the  steam  shut  up  in  the  cylinder. 
In  the  steam  turbine  the  heat  process  is  analogous,  except  that 

*  Knoblauch  and  Jakob,  Zeit.  Verein  deutscher  Ingenieure,  Jan.  5,  1907,  and 
an  article  by  the  author  in  Mechanical  Engineer  (London),  Aug.  24,  1907. 


14  THE   STEAM   TURBINE 

the  flow  of  the  steam,  instead  of  being  intermittent,  is  continuous. 
Steam  is  continually  pushed  into  the  nozzles,  or  similar  steam 
passages,  and  expanding,  expends  its  internal  energy  in  producing 
velocity.  Vanes  or  blades,  'fixed  to  a  rotating  wheel,  are  placed 
near  the  nozzles  so  that  the  jets  of  steam  are  directed  against 
them.  These  blades  or  vanes  thus  set  in  motion  move  the  wheel 
and  with  it  the  shaft  which  transmits  the  power. 

Theoretically  the  work  from  expanding  steam  behind  a  piston 
is  exactly  the  same  as  that  we  obtain  from  a  nozzle.  The  difference 
is  only  in  the  method  for  making  the  heat  energy  available  for 
doing  work. 

Before  going  farther  with  the  discussion  of  -how  the  steam  tur- 
bine converts  heat  energy  into  work,  the  more  familiar  case  of  the 
reciprocating  steam  engine  will  be  considered  briefly,  because  it 
is  assumed  the  reader  is  already  more  or  less  familiar  with  its 
heat  processes.  By  the  static  pressure  in  the  steam  pipes  and  in 
the  boiler  the  steam  is  pushed  into  the  engine  cylinder  and  causes 
the  piston  to  move  up  to  the  point  where  the  supply  of  steam  is 
shut  off.  Then  the  steam  expands,  reducing,  at  the  same  time, 
the  pressure  till  the  piston  has  reached  the  end  of  the  cylinder. 
On  the  return  stroke  the  steam  is  discharged  at  a  nearly  constant 
low  pressure  into  the  atmosphere  or  into  a  condenser.  Now  on 
the  "  working  "  stroke  when  the  steam  is  being  pushed  into  the 
cylinder,*  and  when  it  is  expanding,  the  steam  is  doing  work  at 
the  expense  of  the  heat  energy  put  into  it  by  the  fires  under  the 
boiler.  The  heat  in  a  pound  of  steam  at  a  given  pressure  and 
temperature  represents  a  definite  amount  of  energy.  Expansion 
of  the  steam  in  the  cylinder  after  cut-off  is  accompanied,  therefore, 
with  a  reduction  of  pressure  and  temperature,  and  the  work  done 
is  in  proportion  to  the  heat  energy  lost  by  the  steam.  Thus  heat 
energy  and  work  go  hand  in  hand.  A  loss  to  one  is  a  gain  to  the 

*  Until  the  point  of  cut-off  is  reached,  all  the  time  that  steam  is  being  pushed 
into  the  cylinder  work  is  being  done  at  the  expense  of  the  boiler  pressure.  Actually 
the  pressure  in  the  boiler  is  a  little  lower  after  the  amount  of  steam  required  for  a 
stroke  has  been  taken  out  than  it  was  before..  When,  however,  the  strokes  of 
the  engine  come  in  quick  succession,  the  variation  in  boiler  pressure  is  not  per- 
ceptible. 


THE  ELEMENTARY  THEORY  OF  HEAT 


other.  Fig.  7  shows  a  typical  steam  engine  indicator  card,  repre- 
senting, diagrammatically,  the  heat  relations  that  have  just  been 
discussed.  The  horizontal  scale  of  coordinates  (abscissas)  repre- 
sents volumes,  and  the  vertical  scale  (ordinates)  represents 
pressures.  It  is  obvious  then  that  any  area  included  by  the  lines 
of  this  diagram  represents  work  done  by  the  steam.  In  this 
figure  Pt  and  Vj  represent  initial  pressure  and  volume,  and  P2  and 
v2  the  corresponding  final  conditions,  meaning  the  pressure  and 
volume  at  the  end  of  the  "working"  stroke.  This  diagram  as 
it  applies  to  the  steam  engine  may  be  analyzed  briefly  as  follows: 


FIG.  7.  —  Pressure-Volume  Diagram  Showing  Work  Areas. 

1.  Area  AOiB  is  the  work  done  in  "  pushing"  the  steam  into 
the  cylinder  against  the  resistance  of  the  piston  to  motion. 

2.  Area  I2CB  is  the  work  done  when  the  steam  is  expanding. 

3.  Area  A432C  is  the  work  lost  in  the  heat  energy  discharged 
in  the  exhaust.* 

4.  Area  40123  is  the  net  work  done. 

The  discussion  given  above  is,  of  course,  for  the  ideal  case 
where  the  cylinder  clearance  is  neglected  and  expansion  to 
back-pressure  (P^)  is  complete. 

*  If  an  almost  perfect  vacuum  were  attainable  this  loss  would  be  practically 
negligible.  Actually  with  the  best  condensing  apparatus  it  is  quite  large. 


16  THE   STEAM    TURBINE 

The  same  diagram  (Fig.  7)  can  also  be  used  for  the  analysis  of 
the  work  done  by  steam  expanding  in  the  nozzles  or  similar 
passages  *  of  a  turbine.  The  work  done  in  "  pushing  "  the  steam 
into  the  engine  cylinder  has  its  counterpart  now  in  the  work  done 
by  the  steam  in  entering  the  nozzle,  so  that, 

1.  Area  AOiB  is  the  work  done  in  "  pushing  "  the  steam  out  of 
the  pipes  or  receiving  vessels  into  the  nozzle,  f 

2.  Area  I2CB  is  the  work  done  during  expansion  at  the  expense 
of  the  heat  energy,  to  give  velocity  to  the  steam. 

3.  Area  A432C  is  work  lost  by  the  steam  in  forcing  its  way 
against  the  external  or  exhaust  pressure. 

4.  Area  40123  is  the  work  done  in  producing  velocity. 

The  work  of  "  pushing  "  the  steam  into  the  nozzle  produces 
initial  velocity^,  or  "  velocity  of  approach."  In  all  practical  steam 
turbine  nozzles  this  initial  velocity,  compared  with  the  final 
velocity  after  expansion,  is  very  small.  For  this  reason,  in  the 
calculations  required  for  the  designing  of  nozzles  and  blades, 
this  initial  velocity  is  usually  neglected.  Practical  designers, 
therefore,  are  interested  only  in  the  heat  energy  of  the  area  123 
and  the  velocity  it  represents.  In  order  to  secure  high  efficiency 
and  low  steam  consumption  the  designer  is  always  striving  to 
make  this  area  as  large  as  possible,  allowing,  of  course,  for  other 
limiting  conditions. 

As  the  result  of  the  comparison  of  the  heat  functions  of  steam 
turbines  and  reciprocating  steam  engines,  \ve  should  observe, 
then,  that  the  heat  energy  in  a  pound  of  steam  available  for  per- 
forming useful  work  is  exactly  the  same  whether  the  steam  goes 
to  the  one  or  to  the  other.  It  follows  then  also  that,  theoretically, 

*  In  some  types  of  turbines  there  are  no  nozzles,  but  instead  stationary  blades 
are  used 'which  are  arranged  to  expand  the  steam  just  as  in  a  nozzle.  In  this 
chapter,  therefore,  where  the  term  "nozzle"  is  used  it  will  be  assumed  to  apply  as 
well  to  stationary  "expanding"  blades. 

t  The  amount  of  this  work,  or  the  area  AOiB,  is  very  small  in  the  case  of  the 
turbine  compared  with  that  in  the  steam  engine. 

V   2 

t  This  initial  velocity,  V0,  is  calculated  from  the  relation  P1v1  =  — —  ,    where 

Pl  and  v±  are  the  initial  pressure  and  volume  of   a  pound  of  steam  and  g  is  the 
acceleration  due  to  gravity  (32.2).     All  velocities  are  in  feet  per  second. 


THE  ELEMENTARY  THEORY  OF  HEAT       17 

the  steam  consumption  for  the  same  conditions  of  temperature  and 
pressure  is  the  same  for  the  turbine  as  for  any  other  form  of 
engine.  Discussion  of  the  merits  of  different  forms  of  steam 
motors  with  only  the  theoretical  viewpoint  in  mind  is,  therefore, 
useless.  Only  the  conditions  in  practice  affecting  the  design  of 
commercial  machines  are  of  any  significance  in  determining  the 
type  of  steam  motor  to  be  used  for  given  conditions  of  service. 

HEAT  THEORY  RELATING  TO  THE  DESIGN  OF  NOZZLES 
AXD  BLADES. 

Diagrams  similar  to  those  made  on  a  steam  engine  indicator 
(Fig.  7),  showing  for  an  engine  stroke  the  conditions  of  pressure 
and  volume  inside  the  cylinder,  are  very  useful  in  the  design  and 
operation  of  reciprocating  steam  engines,  but  they  are  of  very 
little  use  for  work  relating  to  steam  turbines.  In  a  steam  tur- 
bine it  is  not  practicable  to  put  a  measured  amount  of  steam 
through  a  nozzle  "  at  a  time  "  as  the  flow  is  practically  con- 
tinuous. The  pressure- volume  diagram  has,  therefore,  a  very 
limited  application.  Another  kind  of  diagram,  the  details  of 
which  are  somewhat  more  difficult  to  understand,  is  universally 
used  by  steam  turbine  engineers.  In  this  diagram,  which  will 
now  be  described,  any  surface  represents  accurately  to  given 
scales  a  quantity  of  heat.  Absolute  temperatures  (T)  are  the 
ordinates,  and  entropies  *  (0)  are  the  abscissas. 

*  Entropy,  which  Perry  calls  the  "  ghostly  quantity,"  has  no  real  physical 
significance,  so  that  complete  definition  is  not  possible.  If  dQ  is  a  small  amount 
of  heat  added  to  a  body,  and  T  is  the  absolute  temperature  at  which  the  heat  is 

added,  then  the  change  in  entropy  of  that  body  is  -^ ,  or  d<t>  =  -^  •  Entropy  of 
saturated  steam  above  the  entropy  of  water  at  the  freezing  point  (32°  F.)  is  easily 
calculated.  For  saturated  steam  at  any  pressure,  then,  </>  =  ^r  +  »  (or  0),  where  x 

is  the  quality  of  the  steam,  r  is  the  latent  heat  of  evaporation  or  "  heat  of  vapor- 
ization," T  is  the  absolute  temperature,  and  n  (or  0)  is  the  entropy  of  the  liquid 
(water).  All  values  of  latent  heat  of  evaporation,  heat  of  the  liquid,  total  heat, 
etc.,  given  in  steam  tables  are  in  heat  units  above  32°  F. 

The  symbols  used  here  are  those  given  in  Peabody's  Steam  and  Entropy  Tables, 
published  by  John  Wiley  &  Sons,  New  York,  and  in  Marks  and  Davis'  Steam 
Tables  and  Diagrams,  published  by  Longmans,  Green  &  Co.,  New  York. 


i8 


THE   STEAM   TURBINE 


000 


400- 


200 


1.0 
Entropy  (0) 

FIG.  8.  —  A  Simple  Entropy- 
Temperature  Diagram. 


2.0 


Fig.  8  shows  a  simple  heat  diagram  laid  out  with  absolute 
temperature  and  entropy  for  the  coordinates.  Steam  at  a  certain 

condition  of  temperature  and 
entropy  is  represented  here 
by  the  point  A.  Then  if  some 
heat  is  added,  increasing  both 
temperature  and  entropy, 
the  final  condition  is  repre- 
sented by  the  point  B,  and 
the  area  ABCD  represents 
the  heat  added  in  passing 
from  the  condition  at  A  to 
the  condition  at  B.  Such  a 
diagram  is  usually  called 
an  entropy-temperature  dia- 
gram, although  the  name 
heat  diagram  would  prob- 
ably be  more  appropriate,  since  every  area  represents  a  definite 
amount  of  heat. 

Another  entropy-temperature  diagram  is  shown  in  Fig.  9, 
representing  by  the  various  shaded  areas  the  heat  added  to  water 
at  32°  F.  to  completely  vaporize  it  at  the  pressure  Pt.  The  un- 
shaded area  under  the  irregular  curve  AB  represents  the  heat  in  a 
pound  of  water  at  the  freezing  point  (32°  F.  or  492°  in  absolute 
temperature).  The  area  OBCD  is  the  heat  added  to  the  water 
to  bring  it  to  the  temperature  of  vaporization,  or  in  other  words, 
this  last  area  represents  the  heat  of  the  liquid  (q)  given  in  the 
steam  tables  for  the  pressure  Px.  Further  heating  after  vaporiza- 
tion begins  is  at  the  constant  temperature  T1  corresponding  to  the 
pressure  Pw  and  is  represented  by  an  increasing  area  under  line 
CE.  When  "  steaming  "  is  complete,  the  latent  heat,  or  the  heat 
of  vaporization  (r),  is  the  area  DCEF.  If  after  all  the  water  is 
vaporized  more  heat  is  added,  the  steam  becomes  superheated, 
and  the  additional  heat  required  would  be  represented  by  an  area 
to  the  right  of  EF. 

The  use  of  the  entropy-temperature  diagram  in  exhibiting  the 


THE  ELEMENTARY  THEORY  OF  HEAT 


behavior  of  steam  during  expansion  and  the  various  heat  losses 
and  exchanges  in  the  passage  of  steam  through  a  turbine  will  now 
be  discussed  and  illustrated  with  a  practical  example. 


II 2 


400 — 

800- 

300  — 


/Steaming        \ 
T!  and  P, \ 


.5  I'.O  ll6  2.0  Entropy(0) 

FIG.  9.  —  Entropy-Temperature  Diagram  showing  the  Total    Heat  in   a  Pound 
of  Dry  Saturated  Steam  at  the  Temperature  Tr 


400 

800- 

300 — 
200 — 


/Steaming 
T, 


\ 


O     M 


.129         .566  1.528        Entropy  (0) 

FIG.  10.  —  Practical  Example  Illustrated  with  an  Entropy-Temperature  Diagram. 

Fig.  10   illustrates  the  heat  process  going  on  when  feed  water 
is  received  in  the  boilers  of  a  power  plant  at  100°  F.,  is  heated  and 


20  THE  STEAM   TURBINE 

converted  into  steam  at  a  temperature  of  400°  F.,  and  then  loses 
heat  in  doing  work.  When  the  feed  water  first  enters  the  boiler 
its  temperature  must  be  raised  from  100°  to  400°  F.  before  any 
"  steaming  "  begins.  The  heat  added  to  the  liquid  is  the  area 
MNCD.  This  area  represents  the  difference  between  the  heat 
of  the  liquid  of  steam  at  400°  F.  (qc)  and  at  100°  F.  (q«)  and  is 
about  306  B.T.U.  The  horizontal  or  entropy  scale  shows  that 
the  difference  in  entropy  between  water  at  100°  and  400°  F.  is 

about  437-* 

Every  reader  should  understand  how  such  a  diagram  is  con- 
structed and  especially  how  the  curves  are  obtained.  In  this 
case  the  curve  NC  is  constructed  by  plotting  from  the  steam  tables 
the  values  of  the  entropy  of  the  liquid  (usually  marked  with  the 
symbol  n  or  6)  for  a  number  of  different  temperatures  between 
1  00°  and  400°  F. 

If  now  water  at  400°  F.  is  converted  into  steam  at  that  tem- 
perature, the  curve  representing  the  change  is  necessarily  a  con- 
stant temperature  line  and  therefore  a  horizontal,  CE.  Provided 
the  vaporization  has  been  complete,  the  heat  added  in  the 
"  steaming  "  process  is  the  latent  heat  or  heat  of  vaporization 
of  steam  (r)  at  400°  F.,  which  is  approximately  827  B.T.U. 

The  change  in  entropy  during  vaporization  is,  then,  the  heat 
units  added  (827)  divided  by  the  absolute  temperature  at  which 
the  change  occurs  (400  +  460  =  860°  F.  absolute)  or 

r       827 

= 


The  total  entropy  of  steam  completely  vaporized  at  400°  F. 
is,  therefore,  the  sum  of  the  entropy  of  the  liquid  (water)  .566  and 
the  entropy  of  the  steam  .962,  or  1.528.!  To  represent  then  by 
CE  this  final  condition  of  the  steam,  the  point  E  is  plotted  where 
entropy  measured  on  the  horizontal  scale  is  1.528,  as  shown  in  the 

*  As  actually  determined  from  Marks  and  Davis'  Steam  Tables,  pp.  9  and  15, 
the  difference  in  entropy  is  .5663  —  .1295  or  .4368.  Practically  it  is  impossible  to 
construct  the  scales  in  the  figure  very  accurately. 

t  Entropy,  like  the  total  heat  (#),  and  the  heat  of  the  liquid  (q),  is  measured 
above  the  condition  of  freezing  water  (32°  F.). 


THE  ELEMENTARY  THEORY  OF  HEAT 


21 


figurec*  The  area  MNCEF  represents  then  the  total  heat  added 
to  a  pound  of  feed  water  at  100°  F.  to  produce  steam  at  400°  F., 
and  the  area  OBCEF  represents,  similarly,  the  total  heat  (H  in  the 
steam  tables)  in  a  pound  of  steam  at  400°  F.  above  that  in  water 
at  32°  F. 

•Adiabatic  Expansion  and  Available  Energy.  The  practical 
example  illustrated  by  Fig.  10  (repeated  here)  will  also  be  used 
to  explain  how  the  entropy-temperature  diagram  can  be  used  to 


,566 


1.528        Entropy  (0) 


FIG.  10,  —  Practical    Example   Illustrated  with   an   Entropy-Temperature 

Diagram. 

show  how  much  work  can  be  obtained  by  a  theoretically  perfect 
engine  from  the  adiabatic  expansion  of  a  pound  of  steam.  When 
steam  expands  adiabatically  —  without  a  gain  or  loss  of  heat  — 
its  temperature  falls.  Remembering  that  areas  in  the  entropy- 

*  The  point  E  is  shown  located  on  another  curve  ST,  which  is  determined 
by  plotting  a  series  01  points  calculated  the  same  as  E,  but  for  different  pressures. 
If  more  heat  had  been  added  than  was  required  for  vaporization,  the  area  DCEF 
would  have  been  larger  and  E  would  have  fallen  to  the  right  of  ST,  indicating 
by  its  position  that  the  steam  had  been  superheated.  The  curve  ST  is  therefore 
a  "boundary  line"  between  the  saturated  and  superheated  conditions.  This 
curve  can  also  be  plotted  from  the  values  obtained  from  a  table  of  the  entropy  of 
dry  saturated  steam. 


22  THE    STEAM  TURBINE 

temperature  diagram  represent  quantities  of  heat  and  that  in  this 
expansion  there  is  no  exchange  of  heat,  it  is  obvious  that  the  area 
under  a  curve  of  adiabatic  expansion  must  be  zero  ;  and  this  con- 
dition can  be  satisfied  only  by  a  vertical  line  which  is  a  line  of 
constant  entropy.*  For  the  case  in  Fig.  10  the  expansion  curve 
will  lie,  therefore,  along  the  line  EF,  and  if  the  temperature  falls 
to  100°  F.  the  expansion  will  be  from  E  to  G,  and  during  this 
change  some  of  the  steam  has  been  condensed.  If  now  heat  is 
removed  from  this  mixture  of  steam  and  water  till  all  the  steam 
is  reduced  to  the  liquid  state,  but  without  further  lowering  of  the 
temperature,  the  horizontal  line  GN  f  will  represent  the  change  in 
its  condition.  The  quantity  of  heat  absorbed  in  this  last  process 
-  technically  known  as  condensing  the  steam  —  is  represented 
by  the  area  MNGF,  and  the  heat  converted  into  work  is,  therefore, 
the.  area  NCEG;  and  this  is  called  the  available  energy.  By 
means  of  diagrams  like  those  in  the  preceding  figures,  it  will  now 
be  shown  how  the  available  energy  of  dry  saturated  steam  for 
any  given  conditions  can  be  readily  calculated  from  the  data  given 
in  steam  tables. 

Fig.  ii  is  an  entropy-temperature  diagram  representing  dry 
saturated  steam  which  is  expanded  adiabatically  from  an  initial 
temperature  Tx  corresponding  to  a  pressure  Px  to  a  lower  final 

*  Since  in  an  adiabatic  expansion  there  is  no  change  of  entropy,  lines  of  constant 
entropy,  in  practice,  are  often  called  "  adiabatics  ".  It  is  very  rare  in  steam  turbine 
work  that  the  expansion  in  a  nozzle  departs  far  from  the  adiabatic.  For  this  reason 
other  kinds  of  expansion  are  not  mentioned  here. 

t  That  the  steam  might  be  dry  and  saturated,  the  expansion  would  have  had 
to  follow  the  curve  E  T  and  G  would  have  appeared  at  G'. 

The  heat  of  the  liquid,  q,  of  a  pound  of  steam  at  100°  F.  is  represented  by  OBNM, 
and  the  heat  of  vaporization  (r)  is  MNG'F',  so  that  the  total  heat  (q  +  r  or  H) 
is  OBNG'F'.  The  total  heat  of  wet  steam  is  expressed  by  q  +  xr,  where  x  is 
the  quality  or  relative  dryness.  In  the  case  of  this  adiabatic  expansion,  then, 
q  is  as  before  OBNM  and  xr  is  MNGF.  It  is  obvious  also  that  the  lines  NG 
and  NG'  have  the  same  relation  to  each  other  as  the  areas  under  them,  so  that 

line  NG        area  MNGF        xr  NG 


line  NG'  ~  area  MNG'F'        r  '         NG' 


showing  that  the  quality  of  the  steam  at  any  point,  G,  on  a  constant  temperature 
line  (which  for  saturated  steam  is  also  a  constant  pressure  line)  is  determined  as 
in  this  case  by  the  ratio  of  NG  to  NG'. 


THE  ELEMENTARY  THEORY  OF  HEAT        23 

temperature  T2  corresponding  to  a  pressure  P2.  The  other  initial 
and  final  conditions  of  total /  heat  (H)  and  entropy  (0)  are 
represented  by.  the  same  subscripts  i  and  2.  The  available 
energy  or  the  work  that  can  be  done  by  a  perfect  engine  under 
these  conditions  is  the  area  NCEG.  It  is  now  desired  to  obtain  a 


<f)  „     Entropy 


FIG.  i  r.  —  Entropy-Temperature  Diagram  for  Steam  Initially  Dry  and 
Saturated. 


simple  equation  expressing  this  available  energy  Ea  in  terms  of 
total  heat,  absolute  temperature,  and  entropy.  Explanations  of 
the  preceding  figures  should  make  it  clear  that 

Ht  =  area  OBNCEGF, 

H,  =  area  OBNG'F', 

Ea  =  areas  (OBNCEGF  +  FGG'F')  -  OBNG'F', 

Ea  =  H,  -  H2  +  FGG'F', 

therefore  Ert  =  Ht  -  H,  +  (0,  -  &)  T,*  (i) 

An  application  of  this  equation  will  be  made  at  once  to  deter- 
mine the  heat  energy  available  from  the  adiabatic  expansion  of 

*  It  should  be  observed  that  this  form  is  for  the  case  where  the  steam  is  initially 
dry  and  saturated.  For  the  case  of  superheated  steam  a  slightly  different  form  is 
required  which  is  given  on  page  55. 


24  THE   STEAM  TURBINE 

a  pound  of  dry  saturated  steam  at  an  initial  pressure  of  165 
pounds  per  square  inch  absolute  to  a  final  pressure  of  15  pounds 
per  square  inch  absolute. 

Example.      Pt  =  165         Tt  =   .  .  . 

P2  =     15         T2  =  673.0  from  steam  tables.* 

HI  =  1 1 95.0  from  steam  tables. 

H2  =  1150.7  from  steam  tables. 

<£i  =  1.5615  from  steam  tables. 

02  =  1-7549  from  steam  tables. 

Substituting  these  values  in  equation  (i),  we  have 

E0  =  1195.0  -  1150.7  +  (1.7549  ~  !-56l5)  673-°  =  17446 
B.T.U.  per  pound  of  steam. 

Now  if  in  a  suitable  piece  of  apparatus  like  a  steam  turbine 
nozzle,  all  this  energy  that  is  theoretically  available  could  be 
changed  into  velocity,  then  we  have  by  the  well-known  formula 
in  mechanics,!  for  unit  mass, 

—  =  Ea (foot-pounds)  =  Ea  (B.T.U.)  X  778, 


V  =  \/  778  X  2gEa    =  223.7  VBO,  (2) 

where  V  is  the  velocity  of  the  jet  and  g  is  the  acceleration  due  to 
gravity  (32.2),  both  in  feet  per  second. 

Solving  then  for  the  theoretical  velocity  obtainable  from  the 
available  energy  in  the  practical  example  above, 

V  =  223.7  ^174.46  =  223.7  X  13.20  =  2953  feet  per  second.! 

The  important  condition  assumed  as  the  basis  for  the  determina- 
tion of  equation  (i),  that  the  steam  is  initially  dry  and  saturated, 

*  The  values  of  the  properties  of  steam  given  in  the  exercises  are  taken  from 
Marks  and  Davis'  Steam  Tables  and  Diagrams. 

t  See  Church's  Mechanics  of  Engineering,  p.  672,  or  Jameson's  Applied 
Mechanics  and  Mechanical  Engineering,  vol.  I,  p.  47. 

t  Losses  in  nozzles  are  neglected.  A  carefully  made  nozzle  may  have  practi- 
cally 100  per  cent  efficiency.  For  discussion  of  nozzle  losses  see  pages  49,  50,  and  86. 


THE  ELEMENTARY  THEORY  OF  HEAT 


must  not  be  overlooked  in  its  application.     There  are,  therefore, 
two  other  cases  to  be  considered^ 

(1)  when  the  steam  is  initially  wet, 

(2)  when  the  steam  is  initially  superheated. 

Available  Energy  of  Superheated  and  Wet  Steam.  The  super- 
heated condition  is  >  somewhat  complicated  and  will  not  be  dis- 
cussed at  this  place,  because  it  can  be  worked  out  more  simply 
with  the  aid  of  the  entropy-total  heat  chart  in  the  appendix.-  The 
method  to  be  used  for  this  case  will  be  discussed  therefore  after 
the  use  of  this  chart  has  been  explained.  (See  pages  55  and  57.) 

The  case  of  initially  wet  steam  is,  however,  easily  treated  in  the 
same  way  as  dry  and  saturated  steam.  Fig.  12  is  an  example  of 


r. /    Ti  and  Pi  E  \E 


N/          T2  and 


0=0  <f>x<f>l       02        Entropy 

FIG.  12.  —  Entropy- Temperature  Diagram  for  Initially  Wet  Steam. 

the  case  in  hand.  At  the  initial  pressure  P1;  the  total  heat  of  a 
pound  of  wet  steam  (qx  +  x^)  is  represented  in  this,  diagram  by 
the  area  OBNCP"<£X.  The  initial  quality  of  the  steam  (xj  is  repre- 

CE" 

sented  by  the  ratio  of  the  lines  — —  •     The  available  energy  from 

CE 

adiabatic  expansion  from  the  initial  temperature  T\  (correspond- 
ing to  the  pressure  PJ  to  the  final  temperature  T2  (corresponding 


26  THE  STEAM  TURBINE 

to  the  pressure  P2)  is  the  area  NCE"G".     If  we  call  this  available 
energy  Eou,,  we  have 

Eaw  =  area  OBNCEGF  +  FGG'F'-  OBNG'F'  -  G"E"EG, 

Eaw  =  Ht  -  H2  +  (fa  -  &)  T2  -    (&  -  «  (Tt  -  T2),* 

Eaw  =  Ht  -  H2  +  (fa  -  fa)  T2  -  ^  (i  -  xj  (Tt  -  T2).     (I') 

•M 

The  velocity  corresponding  to  this  energy  is  found  by  substi- 
tution in  equation  (2),  just  as  for  the  case  when  the  steam  was 
initially  dry  and  saturated. 

Example.  Calculations  for  the  velocity  resulting  from  adia- 
batic  expansion  for  the  same  conditions  given  in  the  preceding 
example  on  page  24,  except  that  the  steam  is  initially  5  per 
cent,  wet,  are  given  below. 

P!  =  165  Ibs.  abs.       T!  =  826.0°  F. 
P2  =     15  Ibs.  abs.       T2  =  673.0°  F. 

HI  =  1195.0  B.T.U. 

H2  =  1150.7  B.T.U. 

0i  =  1.5615. 

<&  =  1-7549- 

ri  =  856.8  B.T.U. 

xi  =  i. oo  -  .05  =  .95. 

Eaw  =  1195.0  -  1150.7  +  (i.7549  ~  1-5615)  673-0  - 


X.05  (826.0  -  673-0), 
Eaw  =  166.53  B.T.U. 
V     =223.7  ^Eaw  =  2-23.7  X  12.90  =  2886  feet  per  second. 

*  In  general  terms, 

$  =  *—  4-  0.     Here 

<£t  =  -£  +  QI     because  x  =  i. 


THE  ELEMENTARY  THEORY  OF  HEAT 


2/ 


FIG.  13.  — Example  of  a  Well-designed  Nozzle. 


rffh 


DeLaval  Type. 


Nozzle 

Diaphragm 


Curtis  Type. 

FIG.  14.  —  Examples  of  Standard  Designs  of  Nozzles. 


28  THE   STEAM  TURBINE 

It  is  observed  that  the  theoretical  velocity  is  reduced  from 
2953  to  2886  feet  per  second  by  the  presence  of  moisture  in 
the  steam.  The  percentage  reduction  in  velocity  is,  however, 
only  about  2  per  cent,  while  the  amount  of  moisture  is  5  per 
cent. 

The  shaded  area  NCEG  in  Fig.  n  is  also  known  as  the  the- 
oretical Rankine  cycle  for  the  case  where  the  steam  supplied 
is  initially  dry  saturated.  The  available  energy,  therefore,  as 
given  by  equation  (i)  on  page  23,  multiplied  by  778  gives  the 
maximum  theoretical  foot-pounds  of  work  that  can  be  accom- 
plished with  this  cycle,  neglecting  losses,  from  a  pound  of  dry 
,  steam.  There  are  33,000  X  60  foot-pounds  in  one  horsepower- 
hour,  and  hence  dividing  33,000  X  60  by  Ea  X  778  we  get  the 
theoretical  steam  consumption  ("  water  rate")  of  an  engine  or 
turbine  using  the  ideal  Rankine  cycle  with  steam  initially  dry 
saturated.  Similarly  the  area  NCE"  G"  in  Fig.  12  shows  the 
available  work  for  the  theoretical  Rankine  cycle  when  the  steam 
is  initially  wet,  and  the  theoretical  steam  consumption  of  the 
Rankine  cycle  for  this  case  is  33,000  X  60  divided  by  Eaw*  X  778. 

Fig.  32,  page  57,  shows  also  the  Rankine  cycle  for  steam  ini- 
tially superheated.  Calculation  of  theoretical  steam  consump- 
tion is  similar  to  the  cases  already  explained. 

The  most  important  part  of  the  design  of  a  nozzle  is  the  deter- 
mination of  the  areas  of  the  various  sections  —  especially  the 
smallest  section,  if  the  nozzle  is  of  an  expanding  or  diverging 
type.  Various  forms  of  standard  nozzles  are  shown  in  Figs.  13 
and  14.  In  order  to  calculate  the  areas  of  nozzles  we  must 
know  how  to  determine  the  quantity  of  steam  (flow)  per  unit  of 
time  passing  through  a  unit  area.  It  is  very  essential  that  the 
nozzle  is  well  rounded  on  the  "  entrance  "  side  and  that  sharp 
edges  along  the  path  of  the  steam  are  avoided.  Otherwise  it 
is  not  important  whether  the  shape  of  the  section  is  circular, 
elliptical,  or  rectangular  with  rounded  corners.  Typical 
"  square,"  "  rectangular  "  and  circular  nozzle  sections  used  in 
different  makes  of  commercial  turbines  are  shown  in  Fig.  15. 

*  From  Equation  (i')»  page  26. 


THE  ELEMENTARY  THEORY  OF  HEAT        28a 

Example.  Calculate  the  work  done  in  foot-pounds  by  one 
pound  of  steam  expanding  behind  a  piston  in  a  reciprocating 
engine  for  the  conditions  given  in  the  example  on  page  24. 
(See  discussion  on  page  16.)  Ans.  174.2  X  778  ft.-lbs. 


FIG.  15.  —  Sections  of  Nozzles  Used  in  Commercial  Turbines. 

Example.  If  the  flow  of  steam  at  165  pounds  per  square  inch 
absolute  pressure  from  a  nozzle  with  a  cross-sectional  area  of 
.128  square  foot  is  200  pounds  per  second,  what  is  the  velocity 
of  the  discharging  jet? 

At  the  pressure  stated  steam  has  a  specific  volume  of  2.75 
cubic  feet  per  pound  (from  steam  tables). 

Let  V  =  velocity  of  discharge  (ft.  per  sec.) 

A  =  area  of  nozzle  =  .128  sq.  ft. 
AV  =  volume  discharged  (cu.  ft.  per  sec.) 
.128  V  =  200  X  2.75 

V  =  4297  ft.  per  sec. 

If  turbine  blades  could  be  made  to  transform  all  this  velocity 
into  useful  work,  how  much  horsepower  could  be  transmitted  to 
machinery  from  its  shaft? 

WV2 
Kinetic  energy  of  jet  (ft.-lbs.  of  work  per  sec.)  =  - —  > 

where  W  is  the  weight  in  pounds  of  the  steam  flowing  and  g 
is  the  acceleration  due  to  gravity  (32.2  ft.  per  sec.). 

ft.-lbs.  of  work  per  sec. 
Horsepower  =  - 

550 

200  X  (42Q7)2 
.2  X32.2  X  550 
=  about  104,300  h.p. 


28b  THE   STEAM  TURBINE 

Example.  What  is  the  theoretical  steam  consumption  (water 
rate)  of  the  Rankine  cycle  for  the  conditions  given  in  the  ex- 
ample on  page  24;  that  is,  for  steam  initially  dry  saturated? 

33,000  X  60 

;  Ans.  -^^-  —  • 

174.46  X  778 

Example.  What  is  the  theoretical  steam  consumption  of 
the  Rankine  cycle  for  steam  initially  wet  at  the  conditions 
stated  in  the  example  on  page  26? 

A         33.°c°  X  60  . 
166.53  X  778 


CHAPTER   III.  , 
FLOW   OF   STEAM   AND    NOZZLE  DESIGN. 

/Flow  of  Steam  through  Nozzles.  The  weight  of  steam  dis- 
:harged  through  any  well-designed  nozzle  with  a  rounded  inlet, 
(similar  to  those  illustrated  in  Figs.  13  and  14,  depends  only  on 
the  initial  absolute  pressure  (PJ,  if  the  pressure  against  which 
the  nozzle  discharges  (P2)  does  not  exceed  .58  of  the  initial  pres- 
^sure.  This  important  statement  is  well  illustrated  by  the  follow- 
ing example.  If  steam  at  an  initial  pressure  (PJ  of  100  pounds 
per  square  inch  absolute  is  discharged  from  a  nozzle,  the  weight 
of  steam  flowing  in  a  given  time  is  practically  the  same  for  all  values 
of  the  pressure  against  which  the  steam  is  discharged  (P2)  which 
are  equal  to  or  less  than  58  pounds  per  square  inch  absolute. 

If,  however,  the  final  pressure  is  more  than  .58  of  the  initial, 
the  weight  of  steam  discharged  will  be  less,  nearly  in  proportion 
as  the  difference  between  the  initial  and  final  pressures  is  reduced. 
(See  pages  32  and  33.) 

The  most  satisfactory  and  accurate  formula  for  the  "  constant 
flow  "  condition,  meaning  when  the  final  pressure  is  .58  of  the 
initial  pressure  or  less,  is  the  following,  due  to  Grashof,*  where  F 
is  the  flow  of  steam  |  (initially  dry  saturated)  in  pounds  per 

*  Grashof,  Theoretische  Maschinenlehre,  vol.  i,  iii;  Hiitte  Taschenbuch,  vol.  i, 
page  333.  Grashof  states  the  formula, 


but  the  formula  given  in  equation  (3)  is  accurate  enough  for  all  practical  uses. 

t  Napier's  formula  is  very  commonly  used  by  engineers  and  is  accurate  enough 
for  most  calculations.     It  is  usually  stated  in  the  form 


70 

where  F,  'P,,  and  AQ  have  the  same  significance  as  in  Grashof'  s  formula.     The 
following  formula  is  given  by  Rateau,  who  has  done  some  very  good  theoretical 

29 


30  THE  STEAM  TURBINE 

second,  A0  is  the  area  of  the  smallest  section  of  the  nozzle  in  square 
inches,  and  Px  is  the  initial  absolute  pressure  of  the  steam  in 
pounds  per  square  inch, 

F-^.  (» 

or,  in  terms  of  the  area, 

A»  -  ^rl  •  (3') 

These  formulas  are  for  the  flow  of  steam  initially  dry  and 
saturated.  An  illustration  of  their  applications  is  given  by  the 
following  practical  example. 

Example.  The  area  of  the  smallest  section  (AQ)  of  a  suitably 
designed  nozzle  is  .54  square  inch.  What  is  the  weight  of  the 
flow  (F)  of  dry  saturated  steam  per  second  from  this  nozzle  when 
the  initial  pressure  (PJ  is  135  pounds  per  square  inch  absolute 
and  the  discharge  pressure  (P2)  is  15  pounds  per  square  inch 
absolute  ? 

Here  P2  is  less  than  .58  Pl  and  Grashof's  formula  is  applicable, 
or, 


60 

,.,       .54  X  116.5* 

F  =  -^—         — -     =  i. 049  pounds  per  second. 
60 

When  steam  passes  through  a  series  of  nozzles  one  after  the 
other  as  is  the  case  in  many  types  of  turbines,  the  pressure  is 
reduced  and  the  steam  is  condensed  in  each  nozzle  so  that  it 
becomes  wetter  and  wetter  each  time.  In  the  low-pressure  noz- 
zles of  a  turbine,  therefore,  the  steam  may  be  very  wet  although 

and  practical  work  on  steam  turbines,   but  his  formula  is  too  complicated  for 
convienent  use: 

F  =  .001  40.Pi  [15.26  -  .96  (log  Pl  -f  log.  0703)]. 
Common  or  base  10  logarithms  are  to  be  used  in  this  formula. 

*  A  curve  from  which  values  of  ^-^  can  be  read  is  given  on  page  38  (Fig.  19). 

"\ 

The  flow  (F)  calculated  by  Napier's  formula  for  this  example  is  F  =  — — , 

or  1.041  pounds  per  second. 


NOZZLE  DESIGN  31 

initially  it  was  dry.  Turbines  are  also  sometimes  designed  to 
operate  with  steam  which  is  initially  wet,  and  this  is  usually  the 
case  when  low-pressure  steam  turbines  (see  Chapter  IX)  are 
operated  with  the  exhaust  from  non-condensing  reciprocating 
engines  —  a  practice  which  is  daily  becoming  more  common.  In 
all  these  cases  the  nozzle  area  must  be  corrected  for  the  wetness 
of  the  steam.  For  a  given  nozzle  the  weight  discharged  is,  of 
course,  greater  for  wet  steam  than  for  dry;  but  the  percentage 
increase  in  the  discharge  is  not  nearly  in  proportion  to  the  per- 
centage of  moisture  as  is  often  stated.  The  general  equation  for 
the  theoretic  discharge  (F)  from  a  nozzle  is  in  the  form* 


*  The  general  equation  for  the  theoretic  flow  is 


/  T          -  ft  +  l-i 

F-  A    V     *g*P,     \(P*\k       {S*\  k   \, 

A°  V(*-I)V,LU/    UJ  J 


where  the  symbols  F,  A0,  Plt  and  g  are  used  as  in  equations  (2)  and  (3).  P2  is 
the  pressure  at  any  section  of  the  nozzle,  vl  is  the  volume  of  a  pound  of  steam  at 
the  pressure  Pv  and  k  is  a  constant.  The  flow,  F,  has  its  maximum  value  when 

2  k  +  l 

k 


is  a  maximum.     Differentiating  and  equating  the  first  differential  to  zero  gives 

i 


P*!     * 
PI     U  +  i 


P2  is  now  the  pressure  at  the  smallest  section,  and  writing  for  clearness  PQ  for 
P2,  and  substituting  this  last  equation  in  the  formula  for  flow  (F)  above,  we  have 


F  = 


Now  regardless  of  what  the  final  pressure  may  be,  the  pressure  (P0)  at  the  smallest 
section  of  a  nozzle  (^40)  is  always  nearly  .58  Pl  for  dry  saturated  steam.  Making 
then  in  the  last  equation  P0  =  .58  Pl  and  putting  for  k  Zeuner's  value  of  1.135 
for  dry  saturated  steam,  we  may  write  in  general  terms  the  form  stated  above, 


where  K  is  another  constant.      See    Peabody's   Thermodynamics  of  the  Steam 
Engine,  page  132;  Zeuner's  Theorie  der  Turbinen,  page  268  (Ed.  of  1899). 


32  THE  STEAM  TURBINE 

where  P±  is  the  initial  absolute  pressure  and  vl  is  the  specific 
volume  (cubic  feet  in  a  pound  of  steam  at  the  pressure  P^.  Now, 
neglecting  the  volume  of  the  water  in  wet  steam,  which  is  a  usual 
approximation,  the  volume  of  a  pound  of  steam  is  proportional 
to  the  quality  (X).  For  wet  steam  the  equation  above  becomes 
then 


The  equation  shows,  therefore,  that  the  flow  of  wet  steam  is 
inversely  proportional  to  the  square  root  of  the  quality  (xi). 
Grashof's  equations  can  be  stated  then  more  generally  as 


(4) 


These  equations  become,  of  course,  the  same  as  (3)  and  (3') 
for  the  case  where  xl  =  i. 

Flow  of  Steam  when  the  Final  Pressure  is  more  than  .58  of  the 
Initial  Pressure.  For  this  case  the  discharge  depends  upon  the 
final  pressure  as  well  as  uponrthe  initial.  No  satisfactory  formula 
can  be  given  in  simple  terms,  and  the  flow  is  most  easily  calculated 
with  the  aid  of  the  curve  in  Fig.  16  due  to  Rateau.  This  curve  is 

used  by  determining  first  the  ratio  of   the  final  to  the  initial 

p 
pressure  -*,  and  reading  from  the  curve  the  corresponding  CO- 

PI 

efficient  showing  the  ratio  of  the  required  discharge  to  that  cal- 
culated for  the  given  conditions  by  either  of  the  equations  (3) 
or  (4).  The  coefficient  from  the  curve  times  the  flow  calculated 
from  equations  (3)  or  (4)  is  the  required  result.  Obviously  the 
discharge  for  this  condition  is  always  less  than  the  discharge 
when  the  final  pressure  is  equal  to  or  less  than  .58  of  the 
initial. 

The  actual  design  of  the  nozzles  for  a  commercial  turbine  will 


NOZZLE  DESIGN 


33 


be  taken  up  in  the  next  paragraph;  but  before  this  is  done,  one 
other  equation  used  almost  continually  in  nozzle  and  blade  de- 
signs must  be  explained.  It  is  to  find  the  quality  of  the  steam 
after  an  adiabatic  expansion.  The  initial  quality  of  the  steam  is 
usually  determined  by  the  conditions  in  the  boiler  equipment,  or 


1.0 

Ratio  of  FiuaLto  Initial  PressureJijL 

FIG.  16.  —  Coefficients  of  the  Discharge  of  Steam  when  the  Final  Pressure   is 
Greater  than  .58  of  the  Initial  Pressure. 

is  given  in  the  engineer's  specifications  for  a  new  design,  but  the 
quality  of  the  steam  after  each  expansion  must  be  calculated. 
The  general  equation  for  adiabatic  flow  (constant  entropy  *)  is 


and  solving, 


•*  1 


(5) 


where  the  subscript  i  attached  to  the  symbols  refers  to  the  initial 
condition,  and  the  subscript  2  to  the  final. 

*  See  footnote  on  page  17. 


The  terms  6l  and 


34 


THE  STEAM  TURBINE 


are  the  entropies  of  the  liquid  (water)  at  the  initial  and  final  con- 
ditions, and  the  other  symbols  are  used  as  before. 

To  avoid  the  laborious  calculation  of  equation  (5)  to  determine 
the  quality  after  adiabatic  expansion,  curves  of  steam  quality 
have  been  calculated  and  plotted  on  the  entropy-total  heat  chart 
in  the  appendix.  To  illustrate  the  use  of  these  curves  an  example 
is  given  below. 

Example.  Steam  at  165  pounds  per  square  inch  absolute 
pressure  (Pj),  which  is  4  per  cent,  wet  (^  =  .96),  is  expanded 
adiabatically  in  a  nozzle  to  15  pounds  per  square  inch  absolute 
(P2\  What  is  the  quality  after  expansion? 

Method.  A  point  is  first  located  on  the  chart  where  the  quality 
curve  for  x  =  .96  crosses  the  pressure  line  for  165  pounds  as  shown 
diagranimatically  in  Fig.  17.  A  horizontal  line  of  constant 
entropy  drawn  through  this  point  shows  at  its  intersection  with 


FIG.   17. — Illustrates  the  Use  of  the  Entropy-Total  Heat  Chart  to  Determine 
the  Quality  of  Steam  after  Expansion. 


the  pressure  line  for  15  pounds  the  quality  after  expansion.  In 
this  case  the  quality  is  .837.  For  practical  designing  to  get  satis- 
factory results  the  quality  should  be  read  to  three  significant 
figures. 

Nozzle  Calculations.  In  the  calculations  to  determine  the  dimen- 
sions of  a  nozzle  it  is  necessary  to  have  given  the  following  data : 

(i)  the  weight  of  steam  that  is  to  be  delivered  through  the 
nozzle  to  develop  the  required  power  in  the  turbine. 


NOZZLE  DESIGN  35 

(2)  the  initial  and  final  pressures  (Px  and  P2)- 

(3)  the  quality  fa)  of  the  steam  supplied. 

With  these  data^40. is  then  calculated  by  substitution  of  these 
quantities  in  equation  (4').  This  is  the  area  at  the  smallest 
section  or  throat  as  shown  in  Fig.  18. 


FIG.  18.  —  A  Typical  Expanding  Nozzle. 

The  area  of  the  nozzle  can  be  determined  by  simple  calculations 
only  at  the  smallest  section  or  throat.  To  determine  the  area  at 
any  other  section  of  the  expanding  portion  between  the  throat 
and  the  mouth  involves  equations  of  the  form  of  those  at  the  bottom 
of  page  31.  It  is  therefore  convenient  to  determine  the  sections 
other  than  the  throat  by  a  proportional  method.  Now  the  areas 
of  different  sections  depend  on  the  following  three  conditions: 

(1)  the  velocity  of  the  steam. 

(2)  the  specific  volume. 

(3)  the  quality  or  dryness. 

The  essential  condition  to  observe  is  that  the  weight  of  steam 
flowing  per  second  is  the  same  at  every  section;  and  for  the  same 
flow  the  areas  are  inversely  proportional  to  the  velocities  at  any 
two  sections  compared,  and  directly  proportional  to  both  the 
specific  volumes  and  the  qualities.  We  may  then  write  the 

equation 

A        V       v        x 

•** x  /%_        *  0    vy        *P    vx        •£  /£\.\ 

•  *9f  -77"  X  ~~~  X  ~~ )  (y) 


Ao 
I 


36  THE    STEAM  TURBINE 

A0  =  area  of  nozzle  at  the  smallest  section  in  square  inches. 
Ax  =  area  of  nozzle  at  any  section  of  expanding  portion  in 

square  inches. 

V0  =  velocity  of  steam  at  the  smallest  section  in  feet  per  second.* 
Vx  =  velocity  of  steam  at  any  section  in  feet  per  second.* 
v0  =  specific  volume  at  the  smallest  section  in  cubic  feet  per 

pound. 

vx  =  specific  volume  at  any  section  in  cubic  feet  per  pound. 
x0  =  quality  of  steam  at  the  smallest  section. 
xx  =  quality  of  steam  at  any  section. 

V        v        x 

The  product  -  ^  X—  -  X  -  when    calculated    for   the   largest 
V*      v0       x0 

section  or  mouth,  is  often  called  the  expansion  ratio  (see  Fig.  21, 
page  41),  and  is  very  nearly  proportional  to  the  ratio  of  the 
initial  to  the  final  pressure. 

An  example  will  now  be  given  to  show  how  the  actual  area  of 
the  nozzles  of  a  commercial  turbine  can  be  calculated. 
Example.     A  test  of  a  De  Laval  turbine  was  as  follows: 
Pressure  in  the  steam-chest  (Pi)  .211.5  pounds  absolute 
Vacuum  referred  to  30-in.  barometer.  .  26.6  in.  mercury 
Moisture  in  steam  .....................  2.2  per  cent. 

Brake  horsepower  ...........................  .  .  333 

Steam  consumption,  per  brake  horsepower- 
hour  as  weighed  ("  wet  ")••••  ........  I5-5I  pounds 

«  Number  of  nozzles  open  ..............  .  ........  ...  8 

In  this  case  P2  is  given  as  2(5.6  inches  vacuum,  which  is  less 
than  2  pounds  absolute  pressure,  and  is  therefore  less  than  .58  PI 
and  formula  (4')  is  applicable,  so  that  the  "  throat  "  area  of  the 
eight  nozzles  is  expressed  by 


*  Since  practically  all  the  loss  in  a  nozzle  occurs  before  the  steam  "  emerges  " 
from  the  throat,  the  same  coefficient  applies  to  both  V0  and  V^  and  cancels  when 
expressed  in  equation  (6).  The  non-expanding  nozzles  shown  on  page  51  are  no 
more  efficient  than  equally  well  made  expanding  nozzles. 


NOZZLE  DESIGN  37 

where  xl  =  .978,  Pt  =  211.5  pounds  per  square  inch  absolute, 
and 

p  =  333__L£iU  =  1L-5  pounds  wet  steam  per  second. 
3600  3600 

~   X 


0 

(211.  5)'97       3600 

^o  =  -333  X-     ?  X  .989  =  .472  square  inches. 
3600 

The  area  of  the  throat  of  each  nozzle  is  therefore  .0590  square 
inches. 

The  value  of  -  -  was  read  from  the  curve  *  of  -   in 

(211.  5)-97  P/7 

Fig.  19. 

The  nozzles  of  most  commercial  types  of  steam  turbines  are 
made  with  straight  sides  as  shown  in  Fig.  18,  so  that  in  addition 
to  the  area  at  the  throat  only  one  other  area  must  be  found  to 
fully  determine  the  expanding  portion.  This  is  obviously  most 
easily  determined  at  the  mouth,  since  the  velocity  must  be  calcu- 
lated from  the  available  energy  for  an  adiabatic  expansion  from 
Pl  =  211.5  pounds  per  square  inch  absolute  to  P2  =  1.67  pounds 
per  square  inch  absolute  (26.6  inches  vacuum).  This  available 
energy  can  be  calculated  by  equation  (i')  for  initially  wet  steam, 
but  the  calculation  is  laborious,  and  instead  the  energy  will  now 
be  read  from  the  entropy-total  heat  chart  in  the  appendix.  The 
point  is  first  located  on  the  chart  where  the  line  for  211.5  pounds 
pressure  crosses  the  .978  steam  quality  line  (estimated).  Read- 
ing the  scale  of  abscissas  at  this  point  we  find  that  the  total  heat 
energy  in  a  pound  of  steam  at  this  condition  is  1181  B.T.U.  By 
following  a  horizontal  line  from  this  point  across  the  chart  as 
indicated  diagrammatically  in  Fig.  20  till  it  intersects  the  pres- 
sure line  corresponding  to  1:67  pounds  (estimated),  the  total  heat 
energy  escaping  with  the  exhaust  steam  after  adiabatic  expansion 

*  The  curve  was  made  in  this  form  to  make  the  final  form  of  the  result  more 
convenient  for  slide-rule  or  cancellation  calculations. 


THE    STEAM  TURBINE 


Initial  Pressure,  Vl  (5  to  50  Lbs.) 
10          15  20          25  30 


40          45 


60         80        400        120        140        160       .180        200        220        240 
JnitiaMIressure,  P^  (50  to  260  Lbs.) 

FIG.  19. — Curves  Showing  Values  of       9~ 

A    \ 


1181  B.T.U. 


874  B.T.U. 


Total  Heat  (H) 


FIG.  20.  —  Illustrates  the  Use  of  the  Entropy-Total  Heat  Chart  for  Deter- 
mining the  Available  Energy  in  a  Pound  of  Steam. 


NOZZLE  DESIGN  39 

as  read  on  the  scale  of  abscissas  is  874  B.T.U.  The  difference 
between  the  two  readings,  or  307'B.T.U.,  is  the  available  energy 
(Ea).  The  quality  at  the  end  of  expansion  (x2)  as  read  from  the 
curves  is  .767.  In  this  way  the  labor  of  calculating  oc2  is  saved. 
From  the  value  of  the  available  energy  due  to  expansion,  Eaw> 
the  velocity  V2  at  the  mouth  of  the  nozzle  is  calculated  by  equation 
(2),  or 

F,  =  223.7  \/Eaw  =  223-. 7  V307  =  3919  feet  per  second. 

In  order  to  determine  the  ratio  of  the  area  at  the  mouth  of  the 
nozzle  (A2)  to  that  at  the  smallest  section  (^40)  by  equation  (6) 
the  velocity  (F0)  and  the  quality  (#0)  *  must  be  determined. 
These  evaluations  are  most  easily  made  in  the  same  way  as  for 
F2  and  x2  by  means  of  the  entropy-total  heat  chart.  Now  the 
available  energy  Eao  corresponding  to  the  velocity  F0,  must  be 
calculated  for  adiabatic  expansion  from 

Pl  =  211.5  pounds  and  xl  =  .978  to 
P0  =  .58  Pit  =  122.7  pounds. 

This  available  energy  is  44  B.T.U.  and  XQ  is  .939.  The 
velocity  Fo  is,  therefore,  223.7  VE^  —  223.7  \/44  =  1483  feet 
per  second.  J 

*  For  steam  initially  dry  and  saturated,  the  quality  after  adiabatic  expansion 

(x2)  for  all  practical  cases  is  very  nearly  expressed,  empirically,  by  the  equation 


and  the  quality  at  the  throat  (#0)  may  be  taken  as  .965  for  all  practical  cases  regard- 
less of  the  initial  and  final  pressures. 

f  It  is  well  established  by  thermodynamic  calculations  and  by  actual  experi- 
ment that  the  pressure  P0  at  the  'smallest  section  of  a  nozzle  is  always  very  nearly 
.58  of  the  initial  pressure  (PJ. 

J  Very  elaborate  curves  of  the  velocities  resulting  from  the  adiabatic  expansion 
of  dry  saturated  steam  have  been  prepared  and  published  in  some  American  books. 
Considering  the  several  stages  in  nearly  all  types  of  turbines,  such  curves  can  be  of 
very  little  use  to  practical  men,  because  the  condition  that  the  steam  admitted  to 
the  nozzles  is  dry  and  saturated  occurs  infrequently.  That  some  of  the  authors 
neglected  to  mark  the  curves  "for  steam  initially  dry  and  saturated"  deserves 
severe  criticism.  The  curves,  as  given,  are  very  misleading,  as  they  are  appar- 
ently intended  for  general  application  for  all  qualities  and  superheats. 


40  THE    STEAM  TURBINE 

In  addition  to  the  values  already  obtained  it  is  only  necessary 
to  get  VQ  and  v^  (the  specific  volumes  of  dry  saturated  steam  at  the 
corresponding  pressures  P0  and  Pt)  to  determine  all  the  terms  in 
the  equation  for  the  expansion  ratio  as  already  given,  and  putting 
now  the  subscript  2  for  x  in  equation  (6)  to  express  the  conditions 
corresponding  to  the  pressure  P2;  then 


or 


x 


—  -r—  X  '-*—*•  X  .01500  =  1.030  square  inches  (area 


3919      3.042       .939 
at  mouth). 

The  author  has  found  as  the  result  of  some  investigations 
regarding  the  design  of  nozzles  that  the  expansion  ratio  (  —  -)of  a 

\AO/ 

properly  designed  nozzle  is  very  nearly  proportional  to  the  ratio  of 
the  initial  pressure  (Px)  to  the  final  pressure  (P2).  The  curve 
shown  in  Fig.  21  has  been  calculated  on  this  basis  for  widely 
different  conditions  but  for  rather  small  expansions,  and  has  been 
found  to  be  accurate  enough  for  practical  purposes  in  designing 
turbines  of  more  than  one  stage.  A  similar  curve  is  now  being 
used  by  the  nozzle  designers  of  one  of  the  large  manufacturing 
companies.  After  the  relations  shown  by  Fig.  21  had  been 
worked  out,  it  was  found  that  Zeuner  had  arrived  at  a  similar 
result  mathematically  after  making  certain  assumptions;*  but 

*  In  Zeuner's  Theorie  der  Turbinen,  page  270,  the  following  equation  is  given 
to  express  the  ratio  of  the  area  at  the  mouth  to  that  at  the  smallest  section  (expan- 
sion ratio): 

A^  =  .1550 

A,  ' 


r  -fir 


•where  the  terms  A  and  P  are  used  as  in  the  equations  above.      There  is  probably 

some  error  in  Zeuner's  assumptions,  because  actually  values  of  —  are  not  quite 

*o 

P., 
constant  for  varying  values  of  ~  ' 


NOZZLE  DESIGN 


Zeuner's  equation  itself  is  practically  useless  on  account  of  being 
too  complicated.* 

Shapes  of  Expanding  Nozzles.     The  inside  walls  of  the  expand- 
ing portion  of  the  nozzle  are  usually  surfaces  with  straight-line 


28 
26 
24 
22 
20 

& 

lie 

•314 

1 

Sis 

^3  10 

>-H 

3  8 

4 
2 
0 

/ 

/ 

/ 

/ 

/ 

/ 

/ 

/ 

/ 

/ 

/ 

/ 

/ 

/ 

/ 

/ 

/ 

t 

/ 

/ 

/ 

/ 

/ 

/ 

i 

/ 

For  Steam  Initially 
Dry  Saturated  or  Wet 

/ 

/ 

/ 

/ 

/ 

— 

*  — 

— 

P.l=. 

58  fPi 

\  \ 

12345 
Expansion  of  Nozzle  1-35* 

FIG.  21.  —  Curve  Showing  the  Approximate  Relation  between  Expansion  Ratio 
of  a  Nozzle  and  the  Ratio  of  the  Initial  to  the  Final  Pressure. 

elements,  meaning  that  in  any  section  of    the  nozzle  along  the 
axis,  like  Fig.   18,  the  inner  walls  are  shown  by  straight  lines. 

*  The  author's  curve  in  Fig.  21  is  expressed  by, 


A  more  accurate  form  for  pressure  ratios  greater  than  25  is  the  following: 


42  THE  STEAM  TURBINE 

Parenty  has  shown  that,  for  the  highest  efficiency,  theoret- 
ically, such  a  section  along  the  axis  should  be  slightly  elliptical 
with  the  focus  in  the  throat,  but  practically  this  shape  shows  no 
advantage  and  is  much  too  difficult  to  construct.  For  making 
nozzles  like  those  in  Curtis  turbines,  where  the  work  is  done 
largely  with  hand  tools,  the  construction  of  even  the  simplest 
form  is  very  expensive  and  the  cost  of  an  elliptical  curvature  is 
practically  prohibitive.  The  shape  to  give  the  best  expansion 
curve  has  been  the  subject  of  investigation  by  various  experi- 
menters.* As  the  practical  results  are  particularly  interesting, 
it  may  be  well  to  describe  briefly  a  typical  form  of  apparatus 
usually  employed  in  these  experiments  as  shown  in  Fig.  22.  The 


FIG.  22.  —  Searching  Tube  Apparatus  for  Determining  the  Pressures  in  Nozzles. 

nozzle  to  be  tested  is  marked  A  in  the  figure.  The  steam  entering 
the  passage  B  discharges  through  the  nozzle  directly  into  the 
exhaust  pipe  E.  A  small  "searching"  tube,  C,  4s  provided 
which  is  sealed  at  one  end  and  has  a  very  small  hole,  D,  a  short 
distance  from  this  end.  The  other  end  of  the  tube  is  attached  to 
a  mercury  column  or  pressure  gauge.  Suitable  means  are  provided 
for  sliding  the  "  searching  "  tube  with  its  pressure  gauge  back  and 

*  The  conditions  of  pressure  and  velocity  of  steam  inside  a  nozzle  are  discussed 
'very  completely  from  the  mathematician's  viewpoint  in  Die  Dampfturbinen  by 
Stodola,  3rd  edition,  pages  42  to  75,  and  in  Zeitschrift  fur  das  Gesamte  Turbinen- 
ivesen,  Aug.  10,  1906,  pages  325-327. 


NOZZLE  DESIGN 


43 


180 
170 
100 
150 
140 
J 130 

I120 
~110 

"100 

590 

380 

2  70 

i  GO 

£  50 
40 
30 
20 
10 
0 


Expansion 


from 


165  Lbs.  per  £q.  In.  Abs. 


forth  so  that  pressures  can  be  observed  in  different  parts  of  the 
nozzle,  corresponding  to  the  position  of  the  hole  D.  From  these 
observations  a  curve  of 
pressures  may  be  made, 
and  from  this,  together 
with  the  data  of  the  weight 
of  steam  passing  per  unit 
of  time,  a  second  curve  may 
be  developed  showing  the 
corresponding  velocities. 
The  curves  in  Figs.  23,  24, 
and  25  are  examples  of  the 
results  obtained  by  this 
method  for  three  very  dif- 
ferent nozzles.  The  nozzle 
shown  at  the  top  of  Fig. 
23  has  curved  lines,  nearly 
elliptical,  for  its  inside 
walls.  A  pressure  curve 
is  shown  beneath  the  sec- 
tion drawing,  in  it:,  true 
relative  position  corre- 
sponding to  points  along 
the  axis  of  the  nozzle. 
Theoretically  this  shape  of  nozzle  approaches  the  ideal  for  an 
adiabatic  expansion.*  Practical  conditions,  however,  as  stated 

*  A  nozzle  with  a  circular  section  (perpendicular  to  the  axis)  has  less  surface 
exposed  to  the  flow  of  steam  than  a  nozzle  of  any  other  form  of  the  same  length 
and  expansion.  For  this  reason  this  form  should  give  minimum  friction  losses. 
In  practice,  however,  this  type  is  not  often  used  when  the  section  at  the  mouth  of 
the  nozzle  is  made  rectangular,  at  least  when  the  nozzles  are  arranged  in  groups 
with  the  mouths  of  the  several  nozzles  close  together.  There  are  obvious  advan- 
tages from  this  last  construction,  as  first  pointed  out  by  Professor  Riedler,  because 
if  the  nozzle  mouths  are  made  rectangular  and  close  together  a  long  continuous 
"band"  of  steam  is  secured  which  is  approximately  homogeneous  and  of  constant 
velocity.  The  flow  from  the  end  nozzles  is,  of  course,  affected  by  excessive  eddying 
and  other  irregularities  just  as.  single  nozzles.  Efficiency  of  the  end  nozzles  is 
therefore  considerably  less  than  that  of  any  of  the  others  in  the  group. 


2  Lbs.  per  S|q.  In.  Abs. 


(26  Ins,  \facuum) 


Distance  along  Axis  of  Nozzle 


FIG.    23.  —  Expansion  Curve  of  a  Nozzle 
with  an  Elliptical  Axial  Section. 


44 


THE  STEAM  TURBINE 


1*0 


130 


before,  make  the  nozzle  shown  in  Fig.  24  with  expanding  straight- 
line  walls  preferable  if  the  throat  and  mouth  areas  are  properly 
designed.  Fig.  24,  however,  is  intended  to  show  primarily  the 
effect  of  using  a  nozzle  for  non-condensing  service,  which  was 
designed  to  be  used  condensing.  For  this  reason  the  expansion 

in  the  nozzle  is  greater 
than  it  should  be  for  the 
pressures  with  which  it 
is  operating;  and  for  this 
reason  the  pressure  inside 
the  nozzle,  as  illustrated  by 
the  curve,  falls  below  the 
exhaust  pressure.  This 
is  called  over-expansion 
or  "  over-compounding  " 
and  is  always  accompa- 
nied fa$  a  loss  in  efficiency. 
In  feet,  as  will  be  shown 
again  later,  the  effect  of 
over-expansion,  or  making 
a  nozzle  too  large  at  the 
mouth,  reduces  nozzle  effi- 
ciency much  more  than  if 
it  is  made  the  same  per- 
centage too  small.  (See 
Fig.  28.)  The  curves 

>  Distance  alang  Axis  of  Nozzle  in  Fig.     24  show  that     the 

FIG.  24.  — Expansion  Curve  of  a  Nozzle  with  pressure     at    the     mouth 
Straight  Walls.  .g   a  ^  lower    than   the 

atmospheric  exhaust,  and  a  partial  vacuum  is  thus  secured 
at  the  blades  opposite  the  nozzles.  When  such  nozzles  are 
operated  non-condensing  there  is  some  gain  from  the  reduc- 
tion of  disk  and  blade  friction  because  the  wheel  and  blades 
revolve  in  a  less  dense  medium;  but  when  considering  also 
the  increased  losses  in  the  nozzle  itself  because  of  over- 
expansion,  there  is  certainly  no  net  gain  over  having  a  nozzle 


g     60 


30 


\.  Atmospheric 


Line 


NOZZLE  DESIGN 


45 


designed  exactly  for  the  expansion  corresponding  to  the  operat- 
ing conditions. 

Fig.  25  is  intended  to  show  an  abnormal  but  interesting  form  of 
nozzle  which  gives  some  idea  of  the  behavior  of  steam  when  the 
expansion  is  not  gradual  and  continuous.  It  was  argued  by  a 
designer  who  made  this  nozzle  that  this  form  should  be  as  efficient 
as  any  other.  It  was 

re$$j$S$S^-V  S1-*  -^ 

x 


his  theory  that  if  the 
•areas  at  the  throat 
and  at  the  .mouth 
were  of  the  right  size, 
the  shape  of  the  walls 
between  was  of  no 
consequence,  and,  in 
fact,  that  the  steam 
of  itself  would  take 
the  correct  passage. 
Thus  by  preventing 
the  steam  particles 
from  touching  the 
walls  the  friction 
losses  in  the  nozzle 
should  be  reduced.  It 
will  be  observed,  how-  ^  _^  ^^ee  along  Axi8  of  Nozz)e 

ever,    from    the    curve    FIG.  25.  — Expansion  Curve  in  an  Abnormal  Nozzle. 

in  the  figure  deter- 
mined from  some  experiments  with  this  nozzle  that  the  pres- 
sure first  drops  abruptly  in  the  throat  to  .58  of  the  initial, 
as  in  any  other  nozzle,  and  then  forms  a  series  of  waves,  from 
which  it  appears  that  the  particles  of  steam  strike  the  walls  and 
rebound,  to  meet  again  at  a  point,  as  at  A,  where  an  increased 
pressure  is  produced,  and  so  on  till  the  mouth  is  reached.  The 
probable  path  of  the  steam  is  shown  by  the  dotted  lines  in  the 
drawing  of  the  nozzle.  These  experiments  show  therefore  that 
the  steam  will  not  take  the  correct  passage  through  a  nozzle  with- 
out the  provision  of  properly  designed  walls  of  gradually  increasing 


180 


1GO 


»140 


1  100 

lao 


40 


46  THE  STEAM  TURBINE 

area  corresponding  to  the  expansion  required.  The  importance 
of  careful  workmanship  in  the  manufacture  of  nozzles  is  therefore 
obvious. 

The  results  shown  by  Fig.  25  bring  up  naturally  the  discussion 
of  the  proper  length  for  a  nozzle,  as  the  one  in  this  figure  was 
obviously  much  too  long. 

Probably  the  best  designers  of  the  Curtis  types  of  turbines  make 
the  length  of  the  nozzle  depend  only  on  the  initial  pressure.  In 
other  words,  the  length  of  a  nozzle  for  150  pounds  per  square  inch 
initial  pressure  is  usually  made  the  same  for  a  given  type  regard- 
less of  the  final  pressure.  And  if  it  happens  that  there  is 
crowding  for  space,  one  or  more  of  the  nozzles  is  sometimes 
made  a  little  shorter  than  the  others. 

Designers  of  De  Laval  nozzles  follow  practically  the  same 
"  elastic  "  method.  The  divergence  of  the  walls  of  non-con- 
densing nozzles  is  about  3  degrees  from  the  axis  of  the  nozzle,  and 
condensing  nozzles  for  high  vacuums  may  have  a  divergence  of 
as  much  as  6  degrees  *  for  the  normal  rated  pressures  of  the 
turbine. 

The  author  has  used  successfully  the  following  empirical 
formula  to  determine  a  suitable  length,  L,  of  the  nozzle  between 
the  throat  and  the  mouth  (in  inches): 


L  =  Vis  A0,  (7) 

where  A0  is  the  area  at  the  throat  in  square  inches. 

The  design  of  the  nozzle  calculated  in  the  example  on  page  36 
can  now  be  completed  with  the  determination  of  its  proper  length, 


L  =  Vi$  X  .059  =  .9  inch. 

The  important  dimensions  of  nozzles  of  circular  section  suitable 

*  According  to  Dr.  O.  Recke,  if  the  total  divergence  of  a  nozzle  is  more  than 
6  degrees,  eddies  will  begin  to  form  in  the  jet.  There  is  no  doubt  that  a  too  rapid 
divergence  produces  a  velocity  loss. 

When  a  number  of  nozzles  intended  for  different  initial  pressures  are  supplied 
for  use  in  the  same  turbine,  the  length  as  determined  by  the  taper  is  usually  made 
to  correspond  to  the  pressure  that  is  to  be  most  used.  Inspection  of  the  De  Laval 
nozzle  in  Fig.  14  shows  that  it  is  necessary  to  make  all  the  nozzles  of  the  same  length 
for  a  given  size  of  De  Laval  turbine,  so  that  the  nozzles  maybe  used  interchangeably. 


NOZZLE  DESIGN  47 

for  this  De  Laval  turbine  tested  by  Dean  &  Main  may  be  tabu- 
lated as  follows: 

Area  at  throat  (A0),  .0590  square  inch.  Diameter  (D0),  .274 
inch. 

Area  at  mouth  (A2),  1.008  square  inches.  Diameter  (D2), 
1.132  inches. 

Length  of  nozzle  (L)  as  determined  by  equation  (7),  .9  inch. 

Length  of  nozzle  assuming  a  divergence  of  12  degrees,  1.9 
inches. 

It  will  be  observed  from  the  last  calculation  that  a  de- 
signer of  De  Lavai  nozzles  would  make  the  length  about  twice 
that  calculated  by  equation  (7).  The  nozzles  of  De  Laval 
turbines  are  made  unusually  long  largely  for  mechanical  rea- 
sons. There  is  probably  very  little  loss  in  this  additional 
length. 

A  nozzle  of  circular  section  suitable  for  these  conditions  is 
shown  at  the  top  of  page  27  (Fig.  13).  It  will  be  observed  that 
a  rounded  entrance  to  the  nozzle  has  been  made,  If  a  well- 
rounded  entrance  is  not  provided  the  rate  of  flow  through  the 
nozzle  may  be  only  50  to  70  per  cent,  (depending  of  course  on  the 
sharpness  of  the  corners)  of  the  normal  flow  calculated  from 
Grashof's  formulas  given  in  equations  (3)  and  (4).  The  effi- 
ciency is  also  very  much  reduced  if  the  steam  is  not  led  to  the 
throat  along  a  surface  of  gradual  curvature.* 

*  Jude  states  that  a  very  large  rounded  inlet  appears  to  "choke"  the  nozzle 
a  little.  He  admits  that  it  gives- maximum  discharge  "but  at  the  expense  of  kinetic 
energy,  that  is,  of  the  kinetic  energy  effective  in  an  axial  direction."  The  results 
of  Rateau's  experiments  seem  to  show,  however,  that  the  efficiency  of  a  convergent 
nozzle  suitably  rounded  is  unity.  If  any  loss  does  result  from  a  rounded  entrance 
which  is  too  large  it  is  probably  of  negligible  amount. .  Some  conclusions  drawn 
from  Rosenhain's  experiments  reported  in  Proc.  Inst.  Civil  Engineers,  vol.  140, 
may  be  of  interest  in  this  connection.  A  series  of  experiments  was  made  with  vari- 
ous nozzles  working  from  20  pounds  to  200  pounds  per  square  inch  gauge  pressure 
with  atmospheric  exhaust.  The  most  efficient  form  of  nozzle  up  to  about  80  pounds 
gauge  pressure  appears  to  be  a  plain  orifice  in  a  thin  plate,  as  measured  by  nozzle 
reaction  (see  page  58),  but  this  does  not  imply  that  such  a  form  is  the  best  nozzle 
for  a  turbine  under  similar  conditions.  With  this  kind  of  orifice  there  is  too  much 
spreading  of  the  jet,  and  the  internal  eddies  and  whirls  are  too  violent  for  useful 
application  at  the  point  where  the  jet  strikes  the  turbine  vanes. 


48 


THE  STEAM  TURBINE 


It  has  been  shown  by  Stodola's  experiments  that  the  difference 
in  pressure  between  the  outer  and  inner  portions  of  the  jet  inside 
a  nozzle  of  approximately  correct  design  are  practically  negligible. 
The  conclusion  is,  therefore,  that  the  jet  always  completely  fills 
the  nozzle,  and  that  there  is  no  "  zonal  formation,"  meaning 
an  outer  zone  moving  at  a  different  velocity  from  the  inner  one, 

although  there  is  certainly  a 
considerable  amount  of.,  fric- 
tional  dragging  of  the  steam 
at  the  surface.     Obviously,  of 
course,  the  statement  does  not 
hold    for  absurdly  diverging 
forms  of  nozzles,  and  in  such 
cases   the    steam    leaves    the 
walls  with   apparently    much 
loss  of  velocity  as  in  the  ex- 
mple  shown  by  Fig.  25. 
Stodola  observed  also  that 
any  nozzle   the    pressure 


r 

£     60 


in 


12345 
Inches  from  the  end  of  the  Nozzle 


usually  falls  in  the  vicinity 
of  the  throat  to  consider- 
ably less  than  the  discharge 

F,G.26_  Experiments  with  an  Expand-  with    a    sudden    rise 

ing  Nozzle  Showing  the  Effect  of  Vary-    r  ' 

ing  the  Final  Pressure.  immediately   after    the     fall. 

This     effect     is     shown     by 

pressure  curves  in  Fig.  26,  plotted  from  Stodola's  data  taken 
in  a  divergent  nozzle  like  the  one  represented  at  the  top  of 
this  figure.  Similar  effects,  only  more  pronounced,  observed  in 
a  straight  non-expanding  nozzle  with  rounded  inlet  are  shown 
in  Fig.  27.  Here  also  a  sudden  drop  below  the  discharge 
occurs;  and,  peculiarly,  the  point  of  depression  progresses  along 
the  axis  of  the  nozzle  as  the  pressure  decreases.  Very  pro- 
nounced oscillations  are  set  up  which  extend  even  into  the 
exhaust  space  for  a  distance  of  about  one  and  a  half  times  the 
length  of  the  nozzle.  The  oscillations  are  apparently  most 
violent  for  the  middle  range  of  pressure,  and  tend  toward  a 


NOZZLE  DESIGN 


49 


1  2 

Inches  from  Mouth 


minimum    when    the    lower    pressure    approaches     a    perfect 
vacuum. 

In  the  divergent  nozzle,  however,  there  appear  to  be  no  internal 
oscillations  of  pressure  after  those  at  the  throat  have  died  out. 

The  size  and  most  likely  also  the  shape 
of  the  external  space  has  a  considerable 
effect  on  these  oscillations  of  pressure. 

Jude  states  in  this  connection  that 
there  is  a  greater  loss  in  velocity,  due 
to  oscillations  or  eddies,  in  a  square  or 
rectangular  nozzle  than  in  a  circular 
one.  Recent  experience  with  nozzles 
of  this  type  does  not  bear  out  this 
statement,  except  in  the  case  probably 
of  square  or  rectangular  nozzles  with 
no  rounding  at  the  edges.  An  efficiency 
of  97  per  cent,  is  not  unusual  for  prop- 
erly designed  square  and  rectangular  FIG.  27.  — Experiments  with 
shaped  nozzles  without  any  "  square  "  a  Non-expanding  Nozzle 

...  .    .          Showing     the     Effect      of 

edges;  and  circular  nozzles  have  certainly     Varying  the  Final  Pressure 
never  given  99  per  cent,  efficiency. 

Under-  and  Over-Expansion.  The  best  efficiency  of  a  nozzle  is 
obtained  when  the  expansion  required  is  that  for  which  the  nozzle 
was  designed,  or  when  the  expansion  ratio  for  the  condition  of  the 
steam  corresponds  with  the  ratio  of  the  areas  of  the  mouth  and 
throat  of  the  nozzle.  A  little  under-expansion  is  far  better,  how- 
ever, than  the  same  amount  of  over-expansion,  meaning  that  a 
nozzle  that  is  too  small  for  the  required  expansion  is  more  efficient 
than  one  that  is  correspondingly  too  large.*  Fig.  28  shows  a 

*  It  is  a  very  good  method,  and  one  often  adopted,  to  design  nozzles  so  that 
at  the  rated  capacity  the  nozzles  under-expand  at  least  10  per  cent.,  and  maybe 
20  per  cent.  The  loss  for  these  conditions  is  insignificant,  and  the  nozzles  can 
be  run  for  a  large  overload  (with  increased  pressures)  in  nearly  all  types  without 
immediately  reducing  the  efficiency  very  much.  This  applies  especially  to  tur- 
bines governed  by  cutting  out  nozzles  in  the  first  stage  (see  page  221)  and  with  no 
control  of  the  nozzles  in  the  other  stages.  Under-expansion  due  to  a  throttling 
governor  is  also  an  important  condition  affecting  the  efficiency  of  nozzles. 


50  THE  STEAM  TURBINE 

% 

curve  representing  average  values  of  nozzle  loss  used  by  various 
American  and  European  manufacturers*  to  determine  discharge 
velocities  from  nozzles  under  the  conditions  of  under-  or  over- 
expansion.  This  curve  will  be  referred  to  again  in  connection 
with  the  design  of  blades  and  is  very  useful  to  the  practical 
designer. 

Non-expanding  Nozzles.     All  the  nozzles  of  Rateau  turbines 
and  usually  also  those  of  the  low-pressure  stages  of  Curtis  turbines 


Nozzle  Velocity  Loss  Due  to 
Under-or  Over-Expansion 

0  10  t*»  05  00  o' 

/ 

/ 

/ 

/ 

/ 

/ 

/ 

\ 

/* 

"^ 

--*. 

s* 

/ 

~-~. 

"-  —  ^, 

—  -». 

•  —  1 

^ 

*^ 

—  " 

'Percentage  Nozzle  is  too  Small 
A$Jttouth  (Under  Expansion) 


Percentage  Nozzle  is  too  Large 
at  Mouth  (Over  Expausion) 


FIG.  28  —  Curve  of  Nozzle  Velocity  Loss. 

are  made  non-expanding;  meaning,  that  they  have  the  same  area 
at  the  throat  as  at  the  mouth.  For  such  conditions  it  has  been 
suggested  that  instead  of  a  series  of  separate  nozzles  in  a  row  a 
single  long  nozzle  might  be  used  of  which  the  sides  were  arcs  of 
circles  corresponding  to  the  inside  and  outside  pitch  diameters 
of  the  blades.  Advantages  would  be  secured  both  on  account  of 
cheapness  of  construction  and  because  a  large  amount  of  friction 
against  the  sides  of  nozzles  would  be  eliminated  by  omitting  a 
number  of  nozzle  walls.  Such  a  construction  has  not  proved 
desirable,  because  by  this  method  no  well-formed  jets  are  secured 
and  the  loss  from  eddies  is  excessive.  The  general  statement 
may  be  made  that  the,  throat  of  a  well-designed  nozzle  should  have 
a  nearly  symmetrical  shape,  as  for  example  a  circle,  a  square,  etc., 
rather  than  such  shapes  as  ellipses  and  long  rectangles.  The 

*  C.  P.   Steinmetz,  Proc.    Am.    .Soc.   Mech.  Engineers,   May,    1908,  page    628. 
A.  Jude,  The  Theory  of  the  Steam  Turbine,  page  39. 


NOZZLE  DESIGN  51 

shape  of  the  mouth  is  not  important.  In  Curtis  turbines  an 
approximately  rectangular  mouth  is  used  because  the  nozzles  are 
placed  close  together  (usually  in  a  nozzle  plate  like  Fig.  114)  in 
order  to  produce  a  continuous  band  of  steam;  and,  of  course,  by 
using  a  section  that  is  rectangular  rather  than  circular  or  elliptical, 
a  band  of  steam  of  more  nearly  uniform  velocity  and  density  is 
secured. 

Fig.  29  shows  a  number  of  designs  of  non-expanding  nozzles 
used  by  Professor  Rateau.     The  length  of  such  nozzles  beyond 


FIG.  29.  —  Rateau  Non-expanding  Nozzles. 

the  throat  is  practically  negligible.  Curtis  non-expanding  nozzles 
are  usually  made  the  same  length  as  if  expanding  and  the  length 
is  determined  by  the  throat  area.  The  Curtis  nozzles  made  in 
Germany  are  a  little  shorter  than  the  length  calculated  by 
formula  (7). 

Materials  for  Nozzles.  Nozzles  for  saturated  or  slightly  super- 
heated steam  are  usually  made  of  bronze.  Gun  metal,  zinc 
alloys,  and  delta  metal  are  also  frequently  used.  All  these  metals, 
have  unusual  resistance  for  erosion  or  corrosion  from  the  use  of 
wet  steam.  Because  of  this  property  as  well  as  for  the  reason 
that  they  are  easily  worked  with  hand  tools  *  they  are  very  suitable 
materials  for  the  manufacture  of  steam  turbine  nozzles.  Super- 
heated steam,  however,  rapidly  erodes  all  these  alloys  and  also 
greatly  reduces  the  tensile  strength.  For  nozzles  to  be  used  with 
highly  superheated  steam,  cast  iron  is  generally  used,  and  except 

*  Nozzles  of  irregular  shapes  are  usually  filed  by  hand  to  the  exact  size. 


52  THE  STEAM  TURBINE 

that  it  corrodes  so  readily  is  a  very  satisfactory  material.  Com- 
mercial copper  (about  98  per  cent.)  is  said  to  have  been  used  with 
a  fair  degree  of  success  with  high  superheats;  but  for  such  con- 
ditions its  tensile  strength  is  very  low.  Steel  and  cupro-nickel 
(8  Cu  +  2  Ni)  are  also  suitable  materials,  and  the  latter  has 
the  advantage  of  being  practically  non-corrodible. 


SUPERHEATED   STEAM. 

In  the  following  pages  the  important  properties  of  superheated 
steam  with  which  the  modern  engineer  must  deal  will  be  briefly 
discussed.  It  is  generally  recognized  that  a  gain  in  steam  economy 
results  from  the  use  of  superheated  steam  in  either  steam  turbines 
or  reciprocating  engines,  but  an  accurate  analysis  of  tests  for  the 
actual  gain  in  economy  of  a  plant  is  very  difficult  because  there 
are  so  many  factors  entering.  The  peculiar  circumstance,  also, 
that  water  can  exist  indefinitely  in  the  liquid  state  in  the  presence 
of  superheated  steam,  makes  conclusions  from  experimental  data 
often  uncertain. 

Flow  of  Superheated  Steam  through  Nozzles.  The  discharge 
of  superheated  steam  from  a  nozzle  is  one  of  the  most  important 
subjects  of  which  the  engineering  profession  generally  has  no 
correct  data.  The  author  has  observed  in  his  practice  again  and 
again  that  the  formulas  ordinarily  given  for  the  flow  of  super- 
heated steam  were  not  correct  and  more  reliable  data  had  to  be 
found.  The  formulas  given  here  were  actually  determined  from 
the  data  of  Lewicki's  experiments  with  a  3o-horsepower  De  Laval 
turbine  *  but  was  later  checked  with  a  great  mass  of  data  in  the 
possession  of  the  General  Electric  Company.  The  precision 
with  which  the  formula  applies  to  Lewicki's  data  is  shown  in  the 
table  given  on  the  next  page. 

A  formula  was  desired  to  express  the  flow  of  superheated  steam 
discharged  from  a  nozzle  in  the  form  of  formula  (3)  for  the  flow 
of  dry  saturated  steam,  together  with  a  suitable  coefficient  to 

*  T^eit.  Verein  deutscher  Ingenieure,  April  4,  1903,  page  494. 
Mitteilungen  uber  Forschungsarbeiten,  Heft  12  (1904),  Zalentafel  25. 


NOZZLE    DESIGN 


53 


correct  for  the  effect  of  superheat.     A  formula  of  this  form  is 
expressed  by 

A  P  97 

(8) 


F  = 


60  (i  4-  .00065  D) 
60  F  (i +  .000650) 


»  or 


(80 


where  F  is  the  weight  in  pounds  of  superheated  steam  discharged 
per  second,  A0  is  the  area  of  the  smallest  section  of  the  nozzle  in 
square  inches,  Pt  is  the  initial  pressure  in  pounds  per  square  inch 
absolute,  and  D  is  the  superheat  in  degrees  Fahrenheit.* 

Lewicki's  data  for  the  tests  f  given  below  were  in  metric  units 
but  are  recorded  here  in  the  corresponding  English  units. 

Initial  pressure  Pt  =  99.25  pounds  per  square  inch  absolute. 

Final  pressure  P2  =  14.6  pounds  per  square  inch  absolute. 


Number  of  Test. 

• 

2 

463.1 
I36.I 

837.0 
1.089 

910.0 

3 

491.9 
164.9 
824.0 
I.I07 

9II.O 

4 

5 

6 

:  7 

8 

Temperature  of  steam,  degrees  F. 
Superheat,  degrees  F.  (D)  
Flow,  pounds  per  hour  (tests)  .  . 
(i  +  .  00065  D)  

386.6 
59-6 
882.0 
1.038 

917.0 

529.7 
2O2.  7 
804.0 
I.I32 

910.0 

592-7 
265.7 
778.0 
I-I73 

914.0 

619.7 
292.7 
776.0 
1.  190 

910.0 

7°3-4 
376-4 
735-° 
1.245 

914.0 

723.6 
396.6 
729.0 
1.258 

916.0 

Flow,  pounds  per  hour,  corrected 
by  formula  (8)   to  equivalent 
flow  of  dry  saturated  steam  .  . 

Volume  of  Superheated  Steam.  Thermodynamic  relations 
show  that  the  flow  of  superheated  steam  is  inversely  proportional 
to  the  square  root  of  the  specific  volume, J  so  that  from  the 
author's  equation  for  the  flow  of  superheated  steam  (8)  the  fol- 
lowing formula  for  the  specific  volume  is  easily  obtained : 


=  (i  +  .00065 


(9) 


*  It  is  stated  that  Mr.  A.  R.  Dodge  has  shown  practically  the  same  results  from 
the  Newport  tests  of  a  Curtis  turbine  reported  by  Mr.  G.  H.  Barrus. 

t  Mitteilungen  uber  Forschungsarbeiten,  Lewicki,  Heft  12  (1904),  Zalentafel  25. 

J  This  relation  is  discussed  by  the  author  in  Mechanical  Engineer  (London), 
Aug.  24,  1907,  page  277,  and  in  the  Harvard  Engineering  Journal,  June,  1907, 
page  36..  Compare  with  Stodola,  Die  Dampfturbinen,  3rd  ed.,  page  9. 


54  THE  STEAM  TURBINE 

where  vs  is  the  volume  in  cubic  feet  of  one  pound  of  superheated 
steam,  v  is  the  volume  in  cubic  feet  of  one  pound  of  dry  saturated 
steam  at  the  same  pressure  as  for  v.s,  and  D  is  the  superheat  in 
Fahrenheit  degrees.* 

Available  Energy  of  Superheated  Steam.  In  the  following 
paragraphs  the  significance  of  the  entropy-temperature  curves  for 
superheated  steam  will  be  explained,  and  it  will  be  shown  also 
how  they  are  to  be  used  to  determine  the  available  energy  and  the 
corresponding  theoretical  velocity  resulting  from  adiabatic  expan- 
sion in  a  nozzle. 

Specific  Heat  of  Superheated  Steam.  In  modern  practice, 
superheated  steam  often  enters  our  calculations  and  a  troublesome 
modification  of  the  entropy  diagram  results.  The  difficulty 
arises  because  the  specific  heat  of  superheated  steam  is  not  at  all 
accurately  known.  The  diagrams  in  the  appendix  are  calculated 
for  the  specific  heat  determinations  by  Knoblauch  and  Jakob,  f 
The  specific  heat  of  steam  varies  with  the  temperature  and  pres- 
sure as  shown  in  Figs.  30  and  31,  giving  values  of  the  mean  and 
the  true  specific  heat  at  constant  pressure  (Cp). 

True  specific  heat  represents  the  ratio  of  the  amount  of  heat  to 
be  added  to  a  given  weight  of  steam  at  some  particular  condition 
Of  temperature  and  pressure  to  raise  the  temperature  one  degree 
to  that  required  to  raise  the  temperature  of  water  at  maximum 
density  one  degree.  The  mean  specific  heat  is  almost  invariably 
used  in  steam  turbine  calculations. 

Entropy  Diagram  of  Superheated  Steam.  The  graphic  repre- 
sentation of  the  heat  added  during  the  superheating  of  steam 
is  easily  accomplished  with  entropy-temperature  diagrams. 
Fig.  32  shows  the  same  diagram  that  represented  dry  saturated 
steam  in  Fig.  12,  with  the  added  area  EHJF  to  show  the  super- 
heating from  the  temperature,  Tt  corresponding  to  the  pressure 

*  This  formula  gives  values  of  specific  volume  representing  a  fair  average 
of  results  obtained  from  the  formulas  of  Zeuner,  Tumlirz,  Knoblauch,  and  Schmidt 
(based  upon  Hirn's  experiments). 

f  Zeit.  Verein  deutscher  Ingenieure,  Jan.  5,  1907.  Values  of  mean  specific 
heat  are  taken  from  Mechanical  Engineer,  July,  1907,  and  Professor  A.M.  Greene's 
paper  in  Proc.  American  Society  of  Mechanical  Engineers,  May,  1907. 


NOZZLE  DESIGN 


55 


Pj  to  the  temperature  of  the  superheated  steam,  T.s.  The  total 
heat  in  a  pound  of  steam  above  the  freezing  point  is  now  repre- 
sented by  the  area  OBCEHJFO.  For  adiabatic  expansion  of 
superheated  steam  at  the  temperature  Ts  and  pressure  Pl  to  a 
pressure  P2  the  available  energy  is  the  area  CEHKL. 

Too  much  calculation  is  involved  in  the  construction  of  entropy 
diagrams  to  make  a  new  diagram  for  every  particular  case  from 


.06 


.42 

200250300350400450500550 


650    700    750 


Temperature  °F 

FIG.  30.  —  Mean  Values  of  Cp  Calculated  by  Integration  from  Knoblauch  and 

Jakob's  Data. 

the  properties  usually  found  in  steam  tables;  but  the  construction 
of  such  diagrams  should  be  understood.  From  the  explanations 
that  have  preceded,  the  construction  of  all  the  lines  except  EH 
should  be  obvious.  This  line  is  obtained  by  calculating  the 
entropy  of  superheated  steam  for  various  values  of  temperature 
from  the  following  well-known  relations: 


or 


log.  —  ?    =  2.3028  CpJlogio  Ta  -  logio  Ti\ 


56 


THE  STEAM  TURBINE 


where  Cpm  is  the  mean  value  taken  from  the  curves  in  Fig.  30  for 
the  temperature  TV 

General  Remarks  Regarding  Nozzles.     Finally  it  may  be  stated 
that  there  is  practically  no  difference  in  the  efficiency  of  the  nozzles 


0.90 


0.85 


0.45 


100J 


200  250  0300 

Temperature    C. 


400 


FIG.  31.  —  Values  of  the  "  True  "  Specific  Heat  of  Superheated  Steam. 

used  in  commercial  turbines  if  they  have  smooth  surfaces  and 
are  properly  designed  for  the  correct  ratio  of  the  area  at  the 
throat  to  that  at  the  mouth,  and  if  the  length  is  not  made  much 
less  than,  nor  more  than  possibly  twice  that  calculated  by 
formula  (7). 


NOZZLE  DESIGN 


57 


Whether  the  nozzle  section  is  throughout  circular,  square,  01 
rectangular  (if  these  last  sections  have  rounded  corners)  the 
efficiency  as  measured  by  the  velocity  will  be  about  96  to  97  per 
cent.,  corresponding  to  an  equivalent  energy  efficiency  of  92  to  94 


f>s    Entropy 
FIG.  32.  —  Entropy-Temperature  Diagram  for  Superheated  Steam. 

per  cent.  Speaking  commercially,  therefore,  it  does  not  seem  to 
be  w,orth  while  to  spend  a  great  deal  of  time  in  the  shops  to  make 
nozzles  very  exactly  to  some  difficult  shape.  Simpler  and  more 
rapid  methods  of  nozzle  construction  should  be  introduced.  In 
some  shops  the  time  of  one  man  for  two  days  is  required  for  the 
hand  labor  alone  on  a  single  nozzle. 

Similarly  to  the  equations  for  available  energy  on  page  23 
for  dry  saturated  steam  and  on  page  26  for  wet  steam,  the 
available  energy  Eas  for  superheated  steam  is  when  the  final  con- 
dition is  "wet": 


When  at  the  final  condition  it  is  superheated,  then 

£„.  =  nl  -  H2  +  c^  (T,  -  ro  -  c_  (T;  -  r,), 

where  other  symbols  are  used  as  before  and  T8  is  initial  tem- 
perature and  T/  is  the  final  temperature  when  superheated. 


CHAPTER  IV. 
STEAM  TURBINE  TYPES  AND  BLADE  DESIGN. 

ALL  the  types  of  both  water  and  steam  turbines  are  commonly 
^divided  into  two  general  classes,  designated  by  the  descriptive 
terms  impulse  and  reaction.  Without  further  explanation,  these 
terms,  as  they  are  used  in  turbine  practice,  would  be  very  mislead- 
ing, because  practically  all  commercial  types  of  steam  turbines 
operate  by  both  the  impulse  and  the  reaction  of  steam.  Long 
usage,  however,  has  determined  the  accepted  meaning  of  these 
terms  and  it  is  useless  now  to  try  to  change  them.  Briefly,  the 


Reaction. 
Force 


.  Impulse 
Force 


FIG.  33.  —  Impulse  of  a  Jet  Exerted  on  a  Flat  Surface. 

physical  phenomena  known  as  impulse  and  reaction  will  first  be 
described,  to  be  followed  by  an  explanation  of  the  technical  sig- 
nificance of  these  terms  as  they  are  used  by  engineers. 

In  all  important  commercial  types  of  steam  turbines  the  blades 

58 


STEAM  TURBINE  TYPES  AND  BLADE  DESIGN 


59 


are  moved  by  both  the  impulse  and  the  reaction  of  impinging 
steam  jets  issuing  from  nozzles  (see  Fig.  2)  or  passages  essentially 
equivalent  to  nozzles.  According  to  the  older  school  of  scientists, 
who  have  handed  down  to  us  the  classification  of  turbines  men- 
tioned above,  an  impulse  is  a  force  acting  in  a  "  forward  "  direc- 
tion, and  a  reaction  is  a  "  backward  "  force,  relative  to  the 
impulse  and  equal  to  it  in  magnitude.  Fig.  33  is  a  simple  con- 
crete illustration  of  both  impulse  and  reaction.  A  suspended 
tank  filled  with  water  is  shown  from  which  a  jet  issues  through  a 
nozzle  and  impinges  upon  a  flat  board  hung  opposite.  As  the 


Reaction 
Force 


Impulse  and 
Reaction  Forces 


FIG.  34.  —  Impulse  of  a  Jet  Exerted  on  a  Curved  Surface. 

result  of  the  pressure  due  to  the  jet,  the  board  will  obviously  move 
to  the  right.  As  the  jet  issues  from  the  nozzle  it  exerts  at  the 
same  time  a  reaction  on  the  tank  causing  it  to  move  to  the  left.* 

*  The  pressure  on  the  walls  of  a  tank  at  any  point  depends  on  the  height  of  the 
water  above  that  point  ("the  head")  and  upon  the  density  of  the  fluid.  When 
a  fluid  escapes  from  an  opening  in  the  tank  there  is  no  resistance  at  that  point  to 
pressure,  and  the  unbalanced  force  exerted  on  the  walls  directly  opposite  will 
tend  to  move  the  tank  in  the  direction  opposite  to  that  of  the  escaping  jet.  The 
greater  the  "head"  and  the  density  the  greater  will  be  the  velocity  of  the  issuing 
fluid  and  the  reaction  on  the  tank, 


6o 


THE  STEAM  TURBINE 


34  is  intended  to  show  the  significance  of  impulse  and 
reaction  as  they  are  used  in  regard  to  turbines.  In  this  case 
water  from  the  tank  impinges  against  the  curved  surface  of  a 
wooden  block,  and  before  it  leaves  this  surface  it  is  turned  back 
upon  itself  through  an  angle  of  180  degrees.  The  block  is  there- 
fore acted  on  by  two  forces  simultaneously,  both  tending  to  move 
it  to  the  right.  When  the  jet  first  strikes  the  surface  of  the  block 


FIG.  35. —  Impulse  Wheel  with  Blades 
of  "  Single  Curvature." 


FIG.  36. — Impulse  Wheel  with 
Blades  of  "Double  Curvature." 


an  impulse  force  tends  to  move.it,  and  when  leaving,  there  is  acting 
in  a  "  backward  "  direction  a  reaction  equal  to  the  impulse.  If  the 
jets  represented  in  the  two  figures  have  the  same  velocity  and 
density,  and  frictional  losses  are  neglected,  the  pressure  on  the 
block  in  Fig.  34  will  be  twice  as  great  as  on  the  board  in  Fig.  33. 
Fig.  35  shows  a  nozzle  and  a  blade  wheel  in  which  the  blades 
have  a  "  single  curvature  "  as  compared  with  the  curved  surface 
in  Fig.  34;  that  is,  the  steam  in  its  passage  through  the  blades  is 
not  "  turned  back  on  itself,"  or  in  other  words,  the  curvature 
of  the  blades  is  less  than  90  degrees.  If  this  wheel  were  held 


STEAM   TURBINE   TYPES  AND -BLADE   DESIGN 


6 1 


stationary  so  that  the  blades  could  not  move,  the  steam  would 
leave  them  in  a  direction  nearly  parallel  to  the  shaft.  The  only 
force,  therefore,  that  is  effective  for  moving  the  blades  is  the 
impulse. 

Fig.  36,  on  the  other  hand,  shows  blades  with  nearly  180  degrees 
curvature  which  turn  the  steam  back  on  itself  on  leaving.  The 
wheel  is  thus  moved  first  by  the  impulse  force  of  the  steam  exerted 
on  the  blades  in  the'direction  of  flow,  and  then  by  its  reaction.  A 
blade  turning  the  steam  through  less  than  90  degrees  like  the  one 
in  Fig.  35  will  exert  only  about  half  as  much  pressure  as  one  turn- 
ing the  steam  through  nearly 
1 80  degrees  like  the  one  in 
Fig.  36. 

A  turbine  wheel  which 
would  be  called  a  reaction 
type  is  shown  in  Fig.  37.  It 
differs  from  the  one  in  Fig.  36 
chiefly  in  the  blade  section  B, 
shown  at  the  top  of  the  draw- 
ing. In  this  type  the  expan- 
sion of  the  steam  in  the  noz- 
zle is  only  partial,  and  the 
blades  are  made  so  that  part 
of  the  expansion  occurs  in 
them.  In  the  types  shown 
in  Figs.  35  and  36,  on  the 
other  hand,  all  the  expansion 
is  in  the  nozzles,  with  no  ex- 
pansion at  all  in  the  blades.* 


FIG.  37.  — Simple  Reaction  Wheel. 


The  amount  of  expansion  of  the  steam  in  the  blades  marks, 
therefore,  the  essential  difference  between  the  two  important 
types  of  steam  turbines  illustrated  by  Figs.  36  and  37.  In  im- 
pulse turbines  there  is  no  expansion  in  the  blades,  while  in 

*  The  turbine  wheel  illustrated  in  Fig.  37  is  not,  however,  typical  of  commercial 
"  reaction  "  types  in  which  there  are  often  a  hundred  or  more  pressure  stages. 


62  THE   STEAM  TURBINE 

reaction  turbines  "  expanding  "  blades  are  used,  with  the  result 
that  some  of  the  kinetic  energy  of  the  steam  is  changed  to 
velocity  in  flowing  through  them. 

From  the  explanation  that  has  preceded  it  is  obvious  that  both 
of  the  types  represented  by  the  last  two  figures  operate  by  both 
impulse  and  reaction. 

Impulse  and  Reaction  of  Fluids.  The  kinetic  energy  of  a 
fluid  jet  discharging  from  a  nozzle  may  be  regarded  as  produced 
by  a  constant  impulse  force  I  acting  upon  a  weight  W  of  the 
fluid  discharged  for  one  second.  During  this  second  the  velocity 
has  changed  from  zero  to  V  feet  per  second  and  has  gone  through 

a  space  of  £  V  feet.     The  work  done  by  this  force  in  producing 

y 
the  kinetic  energy  (K  foot-pounds  per  second)  is  I  X  -,  which 

WV2 
is  equal  to  K  or 

IV      WV2 

We  have  then  —  = •  > 

2  2g 

T*      

g    '* 

In  practice  the  principal  distinguishing  feature  of  reaction  turbines  is  the  applica- 
tion of  stationary  blades  for  partially  expanding  the  steam.  The  rest  of  the  ex- 
pansion takes  place  in  the  moving  blades. 

It  is  sometimes  stated,  although  inaccurately,  that  the  angles  of  the  moving 
blades  may  be  used  as  a  criterion  for  distinguishing  the  two  types.  According  to 
these  authorities,  the  moving  blades  of  impulse  turbines  are  symmetrical  like  Fig.  36, 
and  those  of  reaction  turbines  resemble  in  contour  those  of  Fig.  37.  In  many 
cases  the  rule  could  probably  be  applied,  but  there  are  also  many  exceptions.  There 
are  some  blades  made  for  Curtis  turbines  which  are  not  nearly  symmetrical,  and 
no  one  would  call  a  Curtis  turbine  a  reaction  type. 

The  difference  between  impulse  and  reaction  turbines  can  be  very  easily  shown 
experimentally  by  putting  a  pressure  gauge  between  the  nozzle  and  the  wheel.  In 
the  impulse  type,  because  the  expansion  is  completed  in  the  nozzle,  it  will  be  found 
there  is  no  drop  in  the  pressure  of  the  steam  in  passing  through  the  blades;  but  in 
the  reaction  type  the  gauge  will  record  a  higher  pressure  than  that  in  the  casing. 

As  these  words  "  impulse  "  and  "  reaction  "  are  used  at  the  present  time  there 
is  really  little  connection  between  the  usual  meaning  of  the  words  and  the  ideas 
they  are  to  convey  in  regard  to  steam  turbines.  Actually  all  commercial  steam 
turbines  work  by  impulse  and  by  reaction.  A  German  writer  has  used  instead 
of  "  impulse  "  and  "  reaction  "  the  more  accurate  words,  "  gleichdruck  "  and 
"  ungleichdruck,"  meaning  "  equal  pressure  "  and  "  unequ  al  pressure,"  which  to 
the  author  seem  much  more  appropriate. 


STEAM  TURBINE  TYPES  AND  BLADE   DESIGN  63 

In  the  first  principles  of  physics  it  was  learned  that  impulse 
and  reaction  were  "  equal  and  Opposite,"  so  that  if  the  reaction 

WV 

is  represented  by  R  in  pounds,  then  R  =  I  = 

o 

Example.  If  the  vessel  shown  in  Fig.  33  discharges  10  pounds 
of  water  per  second  at  a  velocity  of  322  feet  per  second,  what  is 
the  force  I  (impulse)  pushing  the  wooden  block  away  from  the 
vessel?  Ans.  100  pounds.* 

Also  what  is  the  force  R  (reaction)  pushing  the  vessel  itself 
toward  the  left?  Ans.  100  pounds. 

Example.  If  water  is  discharged  against  flat  blades  of  a 
water  wheel  made  up  of  vanes  similar  to  the  block  shown  in 
Fig.  33  at  the  rate  of  32.2  pounds  per  second  at  a  velocity  of 
200  feet  per  second  and  is  spattered  from  the  wooden  blocks 
with  a  "  residual  "  velocity  (leaving  the  vanes)  of  100  feet  per 
second,  what  horsepower  is  this  water  wheel  capable  of  devel- 
oping? 

Solution.     Calling  the  "  residual  "  velocity  Vz  we  have 

W  (V2  -   F22)         32.2(2002-I002) 

K  =  -  -^  =  ^-          L  =  15,000  ft.-lbs.  per  sec., 

2g  2X32.2 

15,000 
or  — •  =27.27  horsepower. 

55° 

WV2 

The  maximum  theoretical  horsepower  of  the  wheel  is  -  -  > 

2  g  X  550 
if  the  water  is  discharged  at  zero  velocity.     We  have  (in  this 

x      32.2  X  2oo2         20,000 
case)  — ^—  —  =  —     -  ,    or  36.36  h.p. 

'  2X32.2X550        550 

The  efficiency  of  the  (blades  of  this)  water  wheel  is  therefore 

27.27 

T* — J  =  .75  or  75  per  cent. 

36-36 

Example.  Steam  discharges  from  a  nozzle  at  the  rate  of 
3.542  pounds  per  second  with  a  velocity  of  4000  feet  per  second 
against  the  blades  of  a  steam  turbine  and  leaves  them  with  a 

*  It  is  assumed  that  the  water  leaves  the  block  with  practically  no  velocity, 
that  is,  all  the  velocity  is  absorbed  in  producing  the  impulse  force. 


THE  STEAM   TURBINE 

velocity  of  1000  feet  per  second.  Neglecting  frictional  losses, 
what  is  the  maximum  horsepower  that  this  turbine  wheel  can 
develop?  Calculate  the  efficiency  (percentage)  of  the  blades  in 
this  turbine.  Ans.  1500  horsepower;  93.75  per  cent. 

Example.  The  steam  discharging  from  the  blades  of  the 
turbine  wheel  in  the  last  exercise  is  finally  directed  upon  the 
blades  of  a  second  turbine  wheel.  Assuming  there  has  been  no 
loss  of  velocity  in  passing  from  one  turbine  wheel  to  the  other 
and  that  the  steam  leaves  the  second  one  at  100  feet  per  second, 
calculate  the  maximum  horsepower  that  could  be  developed  in 
this  second  turbine  wheel  and  the  efficiency  of  its  blades. 

Ans.  99  horsepower;  99  per  cent. 

Example.     If  we  consider  the  two  turbine  wheels  mentioned 
in  the  two  preceding  exercises  as  combined  in  a  single  turbine, 
what  would  be  the  total  horsepower  of  the  turbine  and  the 
over-all  efficiency  if  frictional  and  other  losses  are  neglected? 
Ans.  1599  horsepower;  99.94  per  cent. 

Suggestion.  The  same  result  could  have  been  obtained  by 
calculating  the  total  kinetic  energy  of  the  combined  wheels, 
using  V  =  4000  feet  per  second,  V%  =  100  feet  per  second  and 
W  =  3.542  pounds  of  steam. 

Example.  Remembering  that  impulse  and  reaction  are  equal 
and  opposite,  what  is  the  force  of  the  reaction  against  the  plate 
supporting  the  nozzle  required  to  give  a  velocity  of  4000  feet 
per  second  to  a  flow  of  3.542  pounds  of  steam  per  second? 

Ans.  440  pounds. 

Suggestion.     Reaction  =  impulse  (/)  =  - 

o 

Example.  The  area  of  a  nozzle  is  .322  square  inch.  How 
many  pounds  of  steam  per  second  having  a  density  of  .144  pound 
per  cubic  foot  must  be  discharged  from  the  nozzle  in  order  to 
exert  a  pressure  of  90  pounds  against  a  plate  suitably  designed 
to  turn  away  the  steam  with  zero  velocity?  Ans.  .966  pound. 

Suggestion.  In  this  case  all  the  velocity  is  absorbed  in  produc- 
ing the  pressure  (impulse)  upon  the  plate. 


STEAM   TURBINE  TYPES  AND   BLADE   DESIGN  6$b 

Substituting  the  values  given  in  the  example  and  substituting 
in  the  equation  for  impulse,  we' have 

,,,     .322  X  V  X  .144 
W  =  3—  *  =  .000322  F, 

144 

T      WV      .000322  F2  T72 

/  =  -  — ^ =  .ooooi  F2. 

£  32-2 

and  since  the  impulse  is  90  pounds;  we  have 

.00001  F2  =  90  pounds 
F2  =  9,000,000 
F  =  3000  feet  per  second. 

Substituting  this  value  of  V  in  the  equation  at  the  top  of  the 
page, 

W  =  .000322  X  3000  =  .966  pound  per  second. 

Example.  Steam  of  the  same  density  as  in  the  preceding 
exercise  discharges  at  the  rate  of  3478  pounds  per  hour  and  pro- 
duces a  reaction  against  the  plate  into  which  the  nozzle  is  in- 
serted of  90  pounds.  What  is  the  velocity  of  discharge? 

Ans.  3000  feet  per  second. 

EXAMPLES  OF  IMPULSE   TURBINES. 

A  simple  impulse  turbine  is  represented  by  diagrammatic  draw- 
ings in  Fig.  38.  In  the  shaded  drawings  in  this  figure,  "  Section 
A  "  is  made  by  a  plane  cutting  one  of  the  blades  and  passing 
through  the  center  of  the  shaft.  The  other  view,  "  Section  B," 
shows  a  section  made  by  a  plane  parallel  to  the  shaft  and  passing 
through  the  center  of  one  of  the  nozzles  in  the  turbine.  In  the 
same  figure,  Curve  I  shows  the  decreasing  pressures  in  the  nozzle 
and  the  constant  pressure  through  the  blades.  Curve  II  shows 
similarly  the  velocity  changes.  In  the  nozzle  the  steam  velocity 
increases  as  the  pressure  falls,  while  in -the  blades  the  velocity 
of  the  steam  is  absorbed  in  moving  the  wheel.  This  simple 
impulse  turbine  represented  by  these  diagrams  is  typical  of  the 
De  Laval  type.  These  turbines  have  always  a  single  set  of 
nozzles  and  one  row  of  blades. 


Velocity  Triangles 

FIG.  38.  —  Diagrams  of  a  Single- 
stage  Impulse  Turbine. 


SECTION  A 


t  1^1  t 

^"SECTION  B 


(63c) 


Velocity  Triangles 
FIG.  39.  —  Diagrams  of  an 
Impulse     Turbine     with 
Two  Velocity   Stages. 


STEAM   TURBINE   TYPES   AND   BLADE   DESIGN  63d 

In  Figs.  38,  39,  40,  and  41  illustrating  the  important  types  of 
steam  turbines,  the  direction  of  the  flow  of  the  steam  is  marked 
by  the  symbol  c/3— » and  the  motion  of  the  blades  by  w— >.  The 
moving  blades  are  shown  by  solid  black  to  distinguish  them 
from  the  stationary  blades,  which  are  indicated  by  cross-hatching. 

A  modification  of  the  simple  impulse  type  is  shown  in  Fig.  39. 
The  drawings  marked  "Section  A"  and  "Section  B"  show  a 
turbine  with  two  moving  blade  wheels  and  a  set  of  stationary 
"  intermediate  "  blades.  The  stationary  blades  are  merely  guides 
for  changing  the  direction  of  the  steam  so  that  it  will  enter 
the  second  set  of  moving  blades  at  a  suitable  angle.  Two  blade 
wheels  are  used  instead  of  one  in  order  to  make  it  possible  to  use 
efficiently  a  lower  peripheral  speed  for  the  moving  blades.  The 
reasons  for  this  statement  will  be  discussed  in  another  part  of 
this  chapter.  The  curves  at  the  top  of  the  figure  show,  graph- 
ically, the  relation  between  pressure  and  velocity.  Curve  III 
shows  the  sudden  fall  of  pressure  in  the  nozzle  and  the  constant 
pressure  through  the  three  rows  of  blades.  Curve  IV  shows  first 
the  rapid  increase  in  velocity  as  the  pressure  falls,  and  then  the 
gradual  loss  of  velocity  in  the  moving  blades  as  it  is  given  up  in 
doing  work.  Velocities  represented  in  Curves  II  and  IV  are 
drawn  approximately  to  the  same  scale.  A  comparison  shows 
that  the  reduction  in  velocity  of  the  steam  in  the  first  wheel  as 
represented  in  Curve  IV  is  only  about  half  that  for  the  single  wheel 
in  Curve  II.  The  arrangement  of  blades  represented  in  Fig.  39 
makes  possible  comparatively  low  blade  speeds  with  initially 
high  steam  velocities.  This  method  of  increasing  the  number  of 
rows  of  blades  is  often  used  with  three  rows  of  moving  blades  and 
two  "  intermediate  "  (stationary)  rows;  and  even  four  rows  of 
moving  blades  have  been  used.  Not  much  advantage,  however, 
has  been  shown  from  the  use  of  the  third  and  fourth  rows  of 
moving  blades,  and  this  construction  has  been  generally  abandoned. 
Turbines  of  this  type  are  often  spoken  of  as  having  velocity  stages, 
the  number  of  velocity  stages  being  the  same  as  the  number  of 
rows  of  moving  blades. 

The  Curtis  turbines,  made  by  the  General  Electric  Company, 


64  THE   STEAM   TURBINE 

are  the  best  examples  of  the  type  illustrated  by  Fig.  39  with  several 
rows  of  blades  following  a  set  of  nozzles.  In  the  latest  designs 
of  the  larger  sizes  of  these  turbines  there  are  two  rows  of  moving 
blades  and  one  set  of  "  intermediate  "  blades  for  each  set  of 
nozzles,  so  that  the  arrangement  shown  in  Fig.  39  is  typical  of 
these  designs.* 

In  Fig.  40  another  distinct  type  of  steam  turbine  is  illustrated. 
The  left-hand  half  of  this  figure  represents  a  single  impulse  wheel 
as  in  Fig.  38  and  the  right-hand  half  is  practically  a  duplicate  of 
that  on  the  left.  In  this  construction  each  of  the  halves  —  a 
single  nozzle  or  set  of  nozzles  with  the  blades  following  —  is  called 
a  pressure  stage,  or  very  commonly  it  is  called  simply  a  stage. 
The  difference  between  the  operation  of  this  turbine  and  the 
single  impulse  wheel  in  Fig.  38  is  best  shown  by  comparing  the 
pressure  and  the  velocity  curves  at  the  top  of  the  two  figures.  In 
Curve  I,  showing  the  pressure  for  the  single  impulse  wheel,  the 
steam  drops  from  the  boiler  pressure  to  that  of  the  exhaust  in  a 
single  nozzle,  that  is,  in  a  single  stage.  In  Curve  V  of  Fig.  40 
there  is  about  equal  reduction  of  pressure  in  each  of  the  two 
nozzles,  and  the  velocity  change,  as  Curve  VI  shows,  is  about 
the  same  for  each  of  the  two  stages.  This  figure  represents, 
diagrammatically,  a  number  of  types  that  are  more  complex. 

It  should  be  mentioned  here  that  there  are  often  two  or  more 
groups  of  nozzles  and  blades,  each  like  Fig.  39,  in  succession  (cf. 
Fig.  119).  Each  of  these  groups  is  then  called  a  stage.  In  other 
words,  the  first  set  of  nozzles  and  all  the  rows  of  blades  up  to  the 
next  nozzle  make  the  first  stage,  and  so  on.  This  last  arrangement 
is  typical  of  the  Curtis  turbines  with  more  than  one  pressure 
stage  and  the  various  Rateau  designs. 

*The  blades  shown  in  "Section  A"  of  Fig.  39  have  the  sahie  height  on  the 
"entrance"  and  "exit  "sides.  It  is,  however,  a  very  common  practice  to  make  the 
"exit"  side  of  the  "intermediate"  blades  of  Curtis  turbines  a  little  higher  than 
the  "  entrance  "  side  so  as  to  increase  the  cross-sectional  area  and  thus  allow  for 
the  lessened  velocity,  due  to  friction  and  eddies,  and  thereby  prevent  "  choking  " 
in  the  blades.  There  is  therefore  a  little  expansion  in  these  blades. 


STEAM   TURBINE   TYPES   AND   BLADE   DESIGN 


Velocity  Triangles 

FIG.  40.  —  Diagrams  of  an  Impulse  Turbine  with 
Two  Pressure  Stages. 


66 


THE   STEAM   TURBINE 


DrumRotor 

SECTION  A 


SECTION  B 


VaJ 


V 

Velocity  Triangles 

FIG.  41.  —  Diagrams  of  a  Three-stage  Reaction 
Turbine. 


STEAM  TURBINE  TYPES  AND  BLADE  DESIGN         6/ 

The  Rateau  turbine  has  from  20  to  40  pressure  stages,  with  a 
set  of  nozzles  and  a  single  blade  wheel  for  each.  The  drop  in 
pressure  is  then,  of  course,  comparatively  small  in  each  stage. 

REACTION  TURBINES. 

The  arrangement  of  blades  in  the  well-known  Parsons  turbine 
is  illustrated  in  Fig.  41.  This  is  the  typical  modern  reaction  tur- 
bine. There  are  no  nozzles.  The-  steam  flows  from  the  boiler 
into  the  "  admission  space"  of  the  turbine  (see  "  Section  A") 
with  practically  no  velocity.  From  this  space  it  enters  the  first  set 
of  stationary  blades,  where  it  expands  and  attains  some  velocity 
as  the  pressure  drops.  Curves  VII  and  VIII  show  the  change  of 
velocity  with  change  of  pressure.  When  the  steam  leaves  the  fixed, 
blades  it  enters  immediately  the  first  set  of  moving  blades.  Here 
it  expands  again;  but  at  the  same  time  some  of  the  velocity  from 
the  expansion  is  taken  away,  or,  in  other  words,  the  velocity  is 
reduced  in  moving  the  blade  wheels.  The  pressure  and  velocity 
curves  show  plainly  what  happens  in  turbines  of  this  type  as  the 
steam  passes  alternately  through  the  fixed  and  moving  blades, 
expanding  in  every  row  till  it  escapes  in  the  exhaust.  There  is 
here  considerable  expansion  in  the  moving  blades,  and  conse- 
quently because  the  pressure  is  not  the  same  on  both  sides  of  these 
blades  it  is  called  a  reaction  turbine.  All  the  other  three  types 
(Figs.  38-40)  are  impulse  turbines,  because  the  pressure  is  practi- 
cally the  same  on  both  sides  of  the  moving  blades. 

We  should  observe  here  that  all  the  possible  simple  combina- 
tions have  been  mentioned  except  the  case  of  expansion  only  in 
the  moving  blades  and  with  no  expansion  in  the  stationary  parts. 
Such  an  arrangement  would  be  feasible  but  has  probably  never 
been  used. 

In  a  reaction  turbine  any  two  rows  of  blades,  the  first  stationary 
and  the  second  moving,  make  a  pressure  stage.  In  a  Parsons 
reaction  turbine  there  are  sometimes  more  than  a  hundred  stages. 

Graphical  Diagrams  of  Steam  Velocities.  A  velocity  diagram 
representing  graphically  the  steam  velocities  in  the  passages  of 
each  of  four  types  of  turbines  shown  in  Figs.  38-41  is  represented 


68 


THE  STEAM  TURBINE 


at  the  bottom  of  each  of  these  figures.  These  diagrams,  in  the 
shape  of  velocity  triangles,  are  represented  here  with  the  nozzles 
and  blades  in  their  proper  order.  In  practical  designing,  how- 
ever, this  pictorial  effect  is  omitted  and  only  the  triangles  are 
drawn.  The  lines  of  these  triangles  show  by  their  lengths  the 
magnitudes  of  the  blade  as  well  as  the  steam  velocities  in  the  tur- 
bine. As  all  of  these  triangles  are  drawn  to  the  same  scale,  they 
show  how  different  the  velocities  are  in  the  four  types.  In  each 
case  the  blade  speed  (Vb)  is  taken  at  about  the  value  that  has  been 
found  by  experience  to  give  the  best  efficiency.  Such  velocity 
diagrams  are  used  by  engineers  for  determining  the  best  relation 
between  the  velocity  of  the  blades  and  the  velocity  of  the  steam. 
In  order  to  interpret  such  diagrams  intelligently  the  significance 
of  absolute  and  relative  velocities  *  of  the  steam  must  be  clearly 

*  This  distinction  between  absolute  and  relative  velocities  should  probably  be 
made  plainer  for  those  who  are  unfamiliar  with  these  terms.  A  thorough  under- 
standing of  what  is  meant  by  absolute  and  relative  velocities  is  very  necessary  to 
work  intelligently  with  the  velocity  diagrams  on  which  the  whole  theory  of  turbine 
practice  depends.  Suppose  a  train  is  just  moving  out  of  a  station  at  the  rate  of  30  feet 
per  second,  and  a  man  standing  in  the  middle  of  the  track  behind  the  train  throws 

a  ball  with  a  velocity  of  40  feet  per 
second  through  the  back  door  of  the 
last  car.  Then  a  passenger  in  the  train 
-  will  see  the  ball  moving  through  the  car 
at  a  velocity  of  only  10  feet  per  second. 
In  this  case  the  velocity  of  the  ball,  or 
40  feet  per  second,  is  its  absolute  velo- 
city with  respect  to  bodies  that  are  not 
moving,  and  10  feet  per  second  is  the 
relative  velocity  of  the  ball  in  the  train. 
In  this  connection  a  slightly  different 
case  should  also  be  considered.  Sup- 
pose now  the  ball  is  thrown  upon  a 
boat  moving  in  a  stream  at  a  velocity  of 
30  feet  per  second  by  a  man  standing 
on  the  bank  at  P  as  represented  in  Fig.  42.  Let  us  assume  the  absolute  velocity, 
or  the  velocity  with  which  the  ball  is  thrown,  as  again  40  feet  per  second,  but  that 
now  the  path  of  the  ball  makes  an  angle  of  20°  with  the  direction  of  the  moving  boat. 
Then  the  relative  velocity  of  the  ball  (Fr)  with  respect  to  the  direction  of  the  boat  is 
shown  graphically  by  a  triangle  of  velocities  ABC  in  the  figure,  where  AC  is  the 
absolute  velocity  (VJ  of  the  ball,  BC  is  the  velocity  of  the  boat  (Vb),  and  AB  is 
the  relative  velocity  (Fr)  of  the  ball  with  respect  to  that  of  the  boat. 


FIG.  42. 


STEAM   TURBINE   TYPES    AND    BLADE    DESIGN  69 

understood.  An  absolute  velocity  of  a  body  is  its  velocity  with 
respect  to  immovable  points  on  the  earth.  A  relative  velocity  is 
its  velocity  with  respect  to  points  that  are  also  moving. 

The  direction  of  the  line  representing  the  velocity  of  the 
steam  relatively  to  the  blades  should  be  such  that  the  lines  of 
flow  of  the  steam  enter  the  blade  tangentially  to  the  conven- 
tionally straight  portion  of  the  back  of  the  blade  (see  Figs.  43, 
49,  and  50).  If  the  backs  of  the  blades  are  made  to  any  other 
angle  there  will  be  losses  due  to  impact  and  eddies. 

EFFICIENCY  OF  THE   BLADES  OF  IMPULSE  TURBINES. 

In  the  velocity  diagram  in  Fig.  38,  the  initial  velocity  of  the 
steam  entering  the  nozzle  is  marked  Vi,  the  velocity  in  the 
throat  is  V0,  and  the  absolute  velocity  of  the  steam  as  it  leaves 
the  nozzle  and  enters  the  blades  is  V2,  making  an  angle  a  with  the 
direction  of  motion  of  the  blades.  The  velocity  of  the  blades 
Vfe,  which  is  the  peripheral  velocity  of  the  wheel,  produces  a 
"  relative  "  velocity  of  the  steam  in  the  blades  V^.  The  angle 
0  shows  then  the  theoretical  "  entrance  "  angle  for  the  blades 
that  the  steam  may  enter  without  loss  of  velocity  due  to  shock 
or  impact.  These  angles  a.  and  0  are  marked  plainly  in  the  draw- 
ing of  "  Section  B."  The  relative  velocity  of  the  steam  leaving 
the  blades  is  represented  by  Vr3.  Often  the  blades  for  impulse 
turbines  are  made  symmetrical,  so  that  the  angle  y  on  the 
"  exit "  side  of  the  blades  is  equal  to  the  angle  /?  on  the  "  en- 
trance "  side.  The  absolute  velocity  of  the  steam  leaving  the 
blades  is  found  by  geometrically  subtracting  again  the  blade 
velocity  V6.  The  velocity  of  the  blades  is  always  subtracted  a 
second  time,  because  the  direction  of  the  steam  has  been  reversed 
in  passing  through  them.  The  steam  is  discharged  with  the 
absolute  velocity  V3,  which  is  called  commonly  the  "  residual " 
velocity. 

Conditions  of  Best  Efficiency.  The  condition  for  the  highest 
efficiency  of  this  simple  turbine  (Fig.  38)  will  now  be  discussed. 
The  same  velocities  represented  at  the  bottom  of  Fig.  38  are 
shown  again  with  the  addition  of  an  enlarged  section  of  a  blade 


7o 


THE   STEAM  TURBINE 


in  Fig.  43.  The  notation  is  the  same  as  in  Figs.  38-41.  V2  and 
V3  *  are  the  absolute  velocities  of  the  steam  entering  and  leaving 
the  blade,  of  which  a  shaded  section  is  shown.  V^  and  Vr3  are 


FIG.  43.  —  Velocity  Triangles  for  an  Impulse  Turbine. 

the  corresponding  relative  velocities  of  the  steam  as  it  passes 
through  the  blade.  Now  the  energy  in  the  steam  is  measured, 
of  course,  in  terms  of  its  absolute  velocity,  and  is  proportional 

to  the  square  of  its  velocity,  f    The  energy,  then,  in  a  pound  of 

yz  y  2 

steam  entering  a  blade  is  —  and  on  leaving  is  —  •     The  energy 

taken  away  by  the  blades  is,  therefore,  —  (V22  —  V32).    Here  g 

2g 

is  the  acceleration  due  to  gravity  (32.2),  and  for  all  practical 
purposes  is  a  constant  value.  Energy  converted  into  work  in  a 

*  Observe  that  V2,  V3,  V4,  etc.,  indicate  absolute  velocities,  and  V&,  Vr3,  VH, 
«tc.,  are  relative  velocities.  This  relation  should  be  of  much  assistance  in  reading 
the  diagrams. 

The  order  in  the  use  of  subscripts  follows  the  method  used  for  the  nozzles  in  the 
preceding  chapters.  The  subscript  i  is  still  used  to  represent  the  initial  condition 
of  the  steam  as  it  enters  the  nozzles  of  an  impulse  turbine  or  the  first  row  of  sta- 
tionary blades  in  a  reaction  turbine,  while  the  subscript  o  is  for  the  condition 
at  the  throat  of  a  nozzle.  The  first  "  discharge  "  velocity  either  from  nozzles  or 
.stationary  blades  is  therefore  represented  by  the  subscript  2. 

t  See  discussion  of  kinetic  energy  and  velocity,  page  24. 


STEAM  TURBINE   TYPES   AND  BLADE   DESIGN  71 

turbine  depends  then,  theoretically,  only  on  the  term  (F22  —  F32). 
This  term  will  have  its  best  value,  of  course,  when  F3  is  made 
as  small  as  possible.  The  best  theoretical  conditions  of  blade 
speed  and  steam  velocity  are  shown  in  the  following  discussion: 
In  practice  it  is  usual  to  have  given  (i)  the  velocity  of  the 
steam  entering  the  blades;  (2)  the  "  nozzle  angle  "  (the  angle  at 
which  the  steam  strikes  the  blades) ;  and  usually  in  impulse  tur- 
bines still  another  condition,  (3)  that  the  entrance  and  exit  angles 
(|8  and  7)  are  equal.  The  velocities  that  must  be  considered  for 
these  conditions  are  shown  in  Fig.  43.  Here  V2  is  the  absolute 
velocity  of  the  steam  entering  the  blades,  the  angle  a  is  the  "  noz- 
zle angle  "  and  shows  the  inclination  of  the  nozzle  to  the  plane 
of  the  turbine  wheel.  V6  is  the  peripheral  velocity  of  the  blades, 
Vr2  and  Vr3  are  the  relative  velocities  of  the  steam  in  the 
blades,  and  V3  is  its  absolute  velocity  leaving  the  blades.  By  the 
conditions  stated,  F2  and  the  angle  a  are  known,  and  we  are  to 
find  the  most  suitable  blade  velocity  (V6).  Also  the  angle  $  is 
equal  to  the  angle  7,  although  the  value  of  neither  of  these  angles 
is  assumed.  The  velocities  V2,  V6,  and  V^  will  form  one  triangle 
of  velocities,  and  still  another  triangle  is  made  with  V6,  Vr3,  and 
V3.  The  corners  of  the  latter  triangle  are  marked  i,  2,  3,  and 
from  the  geometry  of  the  figure  this  triangle  is  obviously  equal  to 
the  triangle  i,  2',  3,  marked  by  cross-hatching.  Now,  if  we  as- 
sume there  is  no  loss  of  velocity  due  to  friction  and  shock  in  the 
blades  then  Frz  =  Fr3,  and  the  triangle  i,  2',  3  can  then  be  in- 
verted, and,  putting  the  point  2'  at  2,  it  can  be  made  to  join  up 
with  the  triangle  o,  2,  3  which  shows  the  initial  velocities  at  the 
upper  end  of  the  blade.  The  base  o,  i  of  the  new  triangle  o,  i,  3 
is  now  equal  to  2  Vb  and  we  can  write,  by  the  "Law  of  Cosines" 
the  equation 

V32  =  V22  +  (2  V6)2  -  2  V2  (2  V6)  COS  a,  (I  i) 

or  F22  -  F32  =  4  V2  Vb  cos  a  -  4  Vb\ 

V22  -  V32  =  4  V6  (V2  cos  a  -  V6).  (12) 

In  this  equation  the  term  (F22  —  F32),  which  is  a  measure  of 
the  energy  taken  away  from  the  steam,  is  greatest  when  4  F6 


72  THE   STEAM  TURBINE 

(V2  cos  a  —  Vb)  has  its  largest  value;  *  or  we  get  the  maximum 
energy  taken  from  the  steam  when 

V6  =  -  V2  cos  a,  (13) 

which  is  the  condition  when  the  line  3,  i,  or  V3,  is  perpendicular  to 
V6,  that  is,  when  the  steam  leaves  the  blade  perpendicular  to  the 
plane  of  the  wheel.  t 

The  condition  for  which  the  last  set  of  equations  has  been 
worked  out  represents  the  usual  conditions  in  practice.  That  is 
the  "  nozzle  angle  "  is  usually  assumed  (about  20  degrees),  and 
the  blade  angles  0  and  7  are  made  equal.  For  this  case  equation 
(12),  above,  represents  the  best  blade  conditions,  with  the  abso- 
lute velocity  of  the  steam  entering  the  blades  (F2)  and  the 
velocity  of  the  blades  (Vb)  as  the  only  variables. 

We  can  express  the  efficiency  of  the  action  of  the  blades  by 
dividing  the  energy  taken  away  in  performing  work  by  the  energy 
represented  by  the  velocity  of  the  entering  steam;  thus, 

Energy  taken  away  for  work,  or  the  actual  work  done  = 
V22-V32 

^g 

Total  energy  in  the  steam,  which  is  a  measure  of  the  total  work 

V  2 
possible  =  —  • 

2g 

Efficiency^    Actual  work  done    =  Vg-V,;     g  =  V 

2 


total  work  possible          2g         2g         V2 

Now,  in  equation  (12)  we  have  for  the  best  conditions, 


*  If  we  make  the  substitution  F22  —  F32  =  y,  Vb  =  x,  K  =  F2  cos  a,  then  for 
equation  (12)  we  'can  write  y  =  4  x  (K  —  x)  =  4  Kx  —  4  x2. 
For  the  maximum  value  of  y, 


x  =  \K,  or  Vb  =  i  F2  cos  a. 

f  Without  the  calculus  demonstration  it  is  obvious  that  Vz2  —  V£  is  largest 
for  given  values  of  Vz,  when  Vs  is  smallest,  and  this  is  when  the  line  3,1  in  the 
triangle  0,1,3  is  shortest;  or,  in  other  words,  when  the  direction  of  Vs  is  perpen- 
dicular to  the  direction  of  Vb- 


STEAM  TURBINE  TYPES  AND  BLADE  DESIGN  73 

Then  substituting  this  in  equation  (14), 


If,  further,  the-"  nozzle  angle  "  a  is  20  degrees,  as  is  so  com- 
mon in  practice,  then 


Efficiency  =       b  (.940  -J--  (16) 


The  only  variable  left  in  this  equation  is  the  ratio  -r^,  and 

V<L 

it  follows  then  that  the  efficiency  of  a  single  row  of  blades  with 
a  given  nozzle  angle  and  equal  entrance  and  exit  angles  for  the 
blades  depends  only  on  the  ratio  of  the  velocity  of  the  blades  to 
the  velocity  of  the  steam  discharged  from  the  nozzle. 

Impulse  Force  Due  to  Stream  Flow  Across  Stationary  Blades. 
In  Fig.  43  a  a  stream  of  fluid  is  shown  impinging  on  a  blade  at 


FIG.  43a.  —  Stream  Lines  in  Turbine  Blade. 

Ar  where  the  direction  of  flow  is  horizontal  and  parallel  to  the 
contour  of  the  tip  of  the  blade.  At  A  the  stream  exerts  an  im- 
pulse I  in  the  direction  of  flow,  and  as  it  leaves  the  blade  it  exerts 
a  reaction  R,  parallel  to  the  direction  of  flow  at  the  other  end  but 
opposite  to  the  initial  direction  of  flow.  The  component  of  R 
in  the  direction  at  which  the  stream  enters  the  blade  (horizontal) 
is  R  cos  p,  where  0  is  the  angle  the  leaving  stream  makes  with 


74  THE  STEAM  TURBINE 

its  initial  direction  (horizontal).  But  since  impulse  is  equal  to 
reaction  (see  page  62),  I  =  R.  Consequently  the  total  pressure 
upon  the  blade  due  to  both  impulse  and  reaction  is 

I  +  R  cos  p  or  I  (i  +  cos  p). 

When  the  stream  flow  has  been  turned  through  180  degrees 
in  its  passage  over  the  blade,  ft  =  o,  cos  0  =  i,  and  the  total 
pressure  is  2  /.  It  has  been  shown  (page  62)  that 

T      WV 
I  =  -  > 

g 

and  therefore  total  pressure  on  the  blade  is 


g   ~~ 

Also  when  0  =  90  degrees,  as  is  approximately  the  case  in 
Fig-  33  >  cos  0  =  0,  and  the  total  pressure  is 

I  _  WV 

g  ' 

Impulse  Force  Due  to  Stream  Flow  Across  Moving  Blades. 

When  velocities  of  blades  are  also  considered  the  impulse  of  the 

WV 

stream  entering  the  moving  blade  as  in  Fig.  43  is  —   *  cos  0. 

Similarly  the  stream  leaves  with  the  relative  velocity  Vr3  of  which 

TTTTT 

the  component  in  the  direction  of  motion  of  the  blade  is  -  -cos 

g 

WV 

of  angle  3,  i,  2',  or  -  -  (V6  —  Vr3  cos  7).    Total  impulse  is, 

g 

therefore, 


cos  0  +  (vb_  Vr3  cos  7). 


g  g 

For  further  demonstration  see  Exercise  6  in  Appendix. 

In  Fig.  44  a  curve  is  shown  which  has  been  calculated  to  repre- 
sent equation  (16)  for  varying  values  of  blade  speed  (Vb)  and 
with  an  initial  steam  velocity  (F2)  of  3,000  feet  per  second.  The 
increase  in  efficiency  with  increased  blade  velocity  should  be 


STEAM  TURBINE   TYPES  AND   BLADE   DESIGN 


75 


observed,  and  that  the  highest  efficiency  is  obtained  when  the 
blade  speed  (Vfr)  is  about  half  the  velocity  of  the  steam  discharged 
from  the  nozzle  (V2).  This  is  a  good  rough-and-ready  rule  to 
remember.  If,  then,  the  steam  velocity  is  2,500  feet  per  second, 
the  peripheral  velocity  of  the  blade  wheel,  for  the  highest  effi- 
ciency, should  be  about  1,250  feet  per  second.  For  mechanical 
reasons  it  is  difficult  to  construct  turbine  wheels  to  run  at  speeds 


IPO 


CO 


20 


0         200       400      600       800      1000     1200     1400     1600      1800     20002200 
Blade  Speed  V^  -  Ft.  per  Sec. 

FIG.  44.  —  Curve  of  Efficiency  of  an  Impulse  Turbine  with  One  Row  of  Blades 
and  a  Nozzle  Angle  of  20  Degrees  for  Varying  Blade  Speeds. 

much  greater  than  500  feet  per  second,  so  that  many  designers 
will  generally  use  low  blade  speeds  to  get  velocities  more  suitable 
for  commercial  application,  knowing  well  that  in  this  respect 
they  are  sacrificing  their  highest  efficiency. 

In  designing  blades  for  turbine  wheels  the  entrance  and  exit 
angles  (0  and  7)  should  always  be  made  as  nearly  as  possible  of 
the  size  determined  by  the  velocity  diagrams.  If  the  angles  are 
made  much  different,  there  is  a  sudden  change  in  the  direction  of 
the  steam  instead  of  a  gradual  change,  with  a  consequent  loss 
due  to  shock  or  impact. 

Efficiency  of  Velocity  Stages.  An  impulse  turbine  with  more 
than  one  row  of  moving  blades  in  a  single  pressure  stage  (veloc- 
ity stage  type)  is  represented-  by  Fig.  39.  The  energy  taken 
away  from  the  steam  for  work,  as  expressed  in  equation  (12),  can 
be  readily  modified  to  suit  this  case.  We  should  have  observed 
that  each  time  steam  passes  through  a  moving  blade  the  blade 
velocity  (Vb)  is  twice  taken  away  (subtracted  geometrically) 


76 


THE    STEAM   TURBINE 


in  the  velocity   diagrams     If   there   are   N   rows   of   moving 
blades, 

V22  -  W  +  2  =  4  NV6  (V2  cos  a  -  NVb).*  (12') 

And  similarly  (compare  with  equation  14,  page  72), 


4NVb(V2cosa-NVb) 
V22 


Efficiency  = 
and  for  a  2o-degree  nozzle, 
Efficiency  = 


COS  a  — 


NV, 

Vo 


d7) 


(18) 


NV, 

V2     V  V2 

Efficiency  of  a  Simple  Impulse  Turbine  for  Given  Blade  Speed. 

In  the  discussion  of  the  maximum  blade  efficiency  of  impulse 
turbines  which  has  preceded,  the  velocity  of  the  steam  entering 
the  blades  was  assumed  to  be  known  and  a  suitable  blade  speed 
was  determined  in  terms  of  the  entrance  and  exit  angles,  which 
were  assumed  to  be  equal.  This  is  the  problem  which  arises 
when  a  single-stage  impulse  turbine  is  to  be  designed  for  given 
initial  and  final  pressures.  When,  however,  an  impulse  turbine 
of  more  than  one  stage  is  to  be  designed  with  a  fixed  blade  speed 
(Vb)  of  say  500  feet  per  second,f  it  is  desirable  to  determine  the. 

*  This  can  be  shown  geometrically  very  easily  by  the  method  illustrated  at  the 
top  of  Fig.  43  which  will  be  here  drawn  for  three  rows  of  moving  blades.    As  in. 


the  other  figures,  F2  is  the  velocity  of  the  steam  entering  the  first  row  of  blades 
and  Vfi  =  Fr3;  then  in  Fig.  45. 

F52  =  V22  +  (6  Vb)2  -  2  V*  X  6  Vb  cosa. 

F22  _  F52  =  I2  Vb  (F2  cos  a  -  3  F6);  and  F22  -  Flv+2  =  4  NVb  (F2  cos  a. 
—  NVb),  if  N  is  the  number  of  rows  of  moving  blades. 

t  Many  manufacturers  have  a  standard  blade  speed  and  all  sizes  of  turbines 
are  designed  for  this  standard.  The  blade  speeds  of  impulse  turbines  vary  from 
350  to  1200  feet  per  second.  The  latter  figure,  it  is  stated,  has  been  used  success- 
fully by  a  European  manufacturer. 


STEAM  TURBINE  TYPES  AND   BLADE   DESIGN  77 

pressure  drop  in  the  first  stage  (and  probably  also  in  the  second 
stage,  depending  on  the  action 'of  the  valve  gear)  to  obtain  the 
highest  efficiency  in  this  stage.  This  is  because  the  best  results 
are  obtained  in  most  types  by  getting  a  larger  proportion  of 
work  from  the  first  stage  than  from  the  other  stages.*  Efficiency, 
therefore,  is  a  more  important  consideration  in  this  stage  than  in 
the  others. 

We  have  thus  obtained  a  very  simple  form  for  calculating  the 
efficiency  of  an  impulse  turbine;  but  it  must  not  be  overlooked 
that  if  the  entrance  and  exit  angles  are  not  equal,  and  in  the  case 
of  velocity  stages  if  the  exit  angle  of  the  stationary  "  interme- 
diate "  blades  is  not  the  same  as  the  angle  at  which  the  steam  is 
discharged  from  the  preceding  blades,  these  formulas  must  be 
considerably  modified  and  the  result  would  not  be  nearly  so 
simple.  It  should  be  observed  also  that  all  losses  from  friction 
and  eddies  have  been  neglected.  These  more  practical  con- 
siderations are  discussed  in  connection  with  the  examples  of 
actual  designs  of  blades  on  pages  85  to  96. 

EFFICIENCY  OF  THE   BLADES  OF  REACTION  TURBINES. 

As  in  the  case  of  the  impulse  turbine,  the  expressions  for  energy 
and  efficiency  will  now  be  derived  for  the  reaction  turbine, 
assuming  again  that  there  are  no  losses  to  be  considered.  We 
must  remember  that  in  the  reaction  turbine  there  are  no  nozzles 
for  expanding  the  steam  but  that  the  expansion  occurs  in  both 
the  stationary  and  the  moving  blades,  so  that  as  the  steam  goes 
through  the  turbine  its  velocity  is  gradually  and  continually 
changing. 

We  shall  first  consider  a  reaction  turbine  (Fig.  46)  with  only 
two  sets  of  blades.  As  there  are  no  nozzles,  the  first  set  is,  of 
course,  made  stationary.  The  steam  expands  in  going  through 

*  The  reason  for  designing  the  first  stage  for  the  largest  amount  of  work  — 
from  25  to  50  per  cent,  more  than  in  any  of  the  other  stages  —  is  most  apparent  in 
turbines  operated  by  "cut-off"  governing  like  the  Curtis  and  Wilkinson  turbines. 
This  method  of  governing  permits  a  constant  standard  pressure  (presumably  that 
giving  the  maximum  efficiency)  in  the  first  stage  at  all  loads,  while  with  fluctuating 
loads  the  pressures  will  vary  considerably  in  the  other  stages. 


78  THE   STEAM  TURBINE 

these  stationary  blades  and  attains  the  velocity  V^  *  when  it 
reaches  the  first  set  of  moving  blades.  The  relative  velocity  with 
which  the  steam  enters  the  moving  blades  is  V^.  Now,  in  these 


FIG.  46.  —  Velocity  Diagrams  for  One  Stage  of  a  Reaction  Turbine. 

blades  the  steam  is  again  expanded,  so  that  just  before  it  leaves 
the  moving  blades  its  relative  velocity  is  Vr3,  which  is  greater  than 
V^.  The  absolute  velocity  at  which  it  is  discharged  from  the 
moving  blades  is  V3,  and  we  have  the  following  energy  relations: 

V  2 

—  —  =  kinetic  energy  developed  in  the  stationary  blades,  or  the 

2g 

kinetic  energy  entering  the  moving  blades. 

V  ,2  —  Vy>2 

—  -  5-  =  kinetic  energy  developed  in  the  moving  blades. 

2g 

V  2 

—  =  kinetic  energy  carried  away  in  the  discharged  steam.^ 

2g 

The  actual  work  done  on  the  moving  blades  is  Wk  =  (kinetic 
energy  of  the  steam  entering  the  moving  blades)  +  (kinetic  en- 
ergy developed  in  the  moving  blades)  —  (kinetic  energy  carried 
away),  or 

V«2  2  —        2  2 

W,  =  IL 


2g  2g  2g 

If  the  steam  had  left  the  moving  blades  with  zero  velocity,  and, 
therefore,  no  energy  had  been  carried  away  in  the  discharged 
steam,  the  energy  available  for  work  would  be 


2g  2g 

*  See  note  at  the  bottom  of  page-So^  regarding  this  notation. 


(20) 


STEAM  TURBINE  TYPES  AND  BLADE   DESIGN 


79 


pffi  .  actual  work  done  (19)      V22+Vr32-V,32-V32    ,. 

Efficiency  =  -  — : — — -. — r  =  — __  9    ..  , — 17-; — .  (21) 

total  work  possible  (20)         V22+Vr32-Vr22 

In  the  same  way  the  efficiency  can  be  calculated  for  any  number 
of  rows  of  blades.  Equation  (21)  expresses  the  efficiency  for 
only  two  rows  of  blades  —  one  stationary  and  one  moving  —  or, 
in  other  words,  for  one  stage.  We  shall  now  obtain  the  efficiency 
for  three  stages,  that  is,  for  six  rows  of  blades.  The  correspond- 
ing velocity  diagram  is  shown  in  Fig.  47. 


,.  47>  _  Velocity  Diagrams  for  Three  Stages  of  a  Reaction  Turbine.   '. 


V  2 

— —  =  kinetic  energy  developed  in  the  first  stationary  blades. 

y  2  __y  2 

— — —  =  kinetic  energy  developed  in  the  first  moving 


blades. 

T-^  = 

2S 
blades. 

rr    2  _ 


kinetic  energy   developed   in   the  second  stationary 

1/2 

^  =  kinetic  energy  developed  in  the  second  moving 


blades. 
V  2 


—£-  =  kinetic  energy  developed  in  the  third  stationary  blades. 


80  THE  STEAM  TURBINE 

Y  .2  _  Y  2 

L  =  kinetic  energy  developed  in  the  third  moving 

blades. 

V  2 

^  —  kinetic  energy  carried  away  in  the  discharged  steam. 

We  observe  here  that  the  velocities  V03  and  V/,3  are  not  lost  but 
represent  velocities  that  can  be  effective  in  the  succeeding  stages. 
For  this  reason  their  energies  do  not  enter  the  discussion  of 
efficiency.  The  actal  work  in  moving  the  blades  is  then 


'-v-n  +rvto!  +  V  -  yi 

2g          J      [_2g  2  g          J 


2g 

2  V    2        V      2-1          V     2 

-  -X£i.  (22) 


K 


g  2g 

Now,  in  designing  a  reaction  turbine  it  is  desirable  to 
assume  that  the  blade  velocities  and  the  corresponding  angles 
of  the  blades  are  the  same  and  that  equal  steam  velocities  are 
developed  in  each  of  the  three  stages,  so  that, 

v»2  =  vh,  =  vC2, 

y      —  Y     —  V 

*  r2    ~       v  n    ~      *  re> 

V*  -  Vr,  =  Vn, 


and  W 

2  g 


3\-2  g    '  2  g          \ 

The  total  energy  in  the  steam  available  for  work  in  this  case  is 

[Y   2       V  2  —  V  21 
T^^TT^J- 

The  efficiency  is  then 

\JJ,  V      2j_V      2    V2    IV      2 

**  K  _    Ya2      f    Vr3 v  r2 3    Va3      ^  (2*} 


Wa  Va22  +  Vr32  -    Fr22 

It  is  clear,  then,  that  in  the  expression  for  efficiency  the  last 
term  in  the  numerator  changes  its  coefficient  with  the  number  of 
stages,  and  we  see  in  what  proportion  the  efficiency  is  increased 
with  the  number  of  stages. 


STEAM  TURBINE  TYPES  AND  BLADE  DESIGN          8l 

PRACTICAL   DESIGNING    OF   BLADES. 

In  designing  blades  for  steam  turbines  we  must  determine  with 
accuracy, 

(1)  The  angles  for  the  edges  of  the  blades. 

(2)  The  radial  height  or  length  of  the  blades. 

From  the  preceding  discussion  of  velocity  diagrams  and  blade 
efficiencies  it  should  be  clear  how  the  best  angles  for  the  edges  are 
obtained.  It  is  first  necessary  to  calculate  the  velocity  resulting 
from  adiabatic  expansion  between  the  limits  of  pressures  in  the 
stage  for  which  the  blades  are  intended.  Then  velocity  diagrams 
must  be  constructed,  varying  the  blade  angles  if  the  blade  speed  is 
assumed  till  the  best  efficiency  is  obtained.  This  will  be  when 
the  steam  leaves  the  last  blades  nearly  at  right  angles  to  the 
plane  of  the  wheel;  that  is,  when  the  absolute  velocity  of  the  steam 
leaving  the  blades  is,  in  the  diagram,  nearly  perpendicular  to  the 
line  showing  the  blade  velocity  (see  page  76). 

Design  of  Blades  for  Impulse  Turbines.  We  shall  continue  with 
the  discussion  of  the  design  of  blades  for  an  impulse  turbine  with 
nozzles  and  with  a  single  row 
of  blades,  assuming  now  that 
the  entrance  and  exit  angles 
(/?  and  y)  have  been  deter- 
mined. We  shall  assume  also 
that  the  total  area  of  the  noz- 
zles at  their  largest  section  has 
been  calculated  as  it  has  been 
explained  on  pages  36  to  41. 

To  avoid  losses  by  eddies, 
nozzles  are  often  arranged  in 
groups  placed  symmetrically 
with  respect  to  the  periphery  FIG.  48.  -  Diagram  Showing  Location  of 

of   the  blade  wheel.      Usually  Nozzles  in  a  Diaphragm. 

the  nozzles  would  be  arranged 

in  two  groups  diametrically  opposite  in  a  circular  plate,  called  a 

diaphragm,  as  in  Fig.  48.     We  shall  assume  that  each  nozzle 


82  THE  STEAM  TURBINE 

group  covers  one-fourth  of  the  circumference  of  the  blade  wheel. 
Then  if  the  blades  in  the  wheel  were  removed  so  that  they 
could  not  obstruct  the  flow  of  steam,  the  area  through  which  the 
steam  can  pass  is  approximately  J  nDh  for  each  nozzle  group, 
where  D  is  the  mean  diameter  of  the  blade  wheel  and  h  is  the 
height  of  the  opening  from  which  the  blades  have  been  removed. 
When,  however,  there  are  blades  on  the  wheel  the  height  h 
must  be  increased,  because  the  effective  area  for  the  passage  of 
steam  is  reduced. 

Fig.  49  shows  two  views  of  a  small  segment  of  a  blade  wheel. 
The  pitch  of  the  blades  is  marked  p  and  the  blade  angle  is  ft.  If 
there  are  no  blades,  the  area  for  the  passage  of  steam  in  a  length  p 
is  approximately  p  X  h.  With  the  blades  in  the  wheel  the  area 
is  only  p  X  h  sin  ft.*  It  follows  then,  when  we  have  blades 
under  the  nozzle  groups,  that  the  effective  area  under  each  group 
is  1  TrDh  sin  ft.  If  we  call  A  the  total  area  of  the  nozzles  at  the 
largest  cross-section  (mouth)  we  can  write 

A  =  \  TrDh  sin  ft  +  \  TrDh  sin  ft. 
A  =  i  TrDh  sin  ft. 

'      2  A  ,       . 

h  =  ^ta7'  (24) 

For  steam  at  very  high  velocity  the  height  of  the  blades  as  calcu- 
lated will  be  too  small  for  practical  working  conditions;  so  that 
blades  less  than  .25  inch  high  are  not  often  made.  This  minimum 
height  is  determined  most  on  account  of  mechanical  difficulties; 
but  steam  leakage  through  the  clearance  outside  the  blades  also 
becomes  excessive  when  very  small  blades  are  used. 

In  impulse  turbines  with  only  a  few  stages  no  effort  is  made  to 
make  use  of  the  velocity,  as  such,  of  the  steam  leaving  the  last 
blades  of  a  stage.  This  means  some  loss ;  and  more  experimental 
work  might  well  be  done  with  the  object  of  showing  how  this  loss 
can  be  turned  to  better  account. 

Fig.  50  shows  how  impulse  turbine  blades  are  laid  out.     The 

*  It  is  assumed  in  such  calculations  that  the  thickness  of  the  edges  of  the  blades 
is  practically  negligible. 


STEAM  TURBINE  TYPES  AND  BLADE  DESIGN 


SECTION  AT  CENTER  LINE 


designer  must  first  decide  how  wide  his  blades  shall  be.  For 
turbines  of  less  than  100  horsepower  the  width  of  the  blades  is. 
often  made  about  i  inch,  increas- 
ing this  dimension  to  about 
1.5  inches  in  turbines  of  1,000 
horsepower.  The  pitch  or  cir- 
cumferential distance  between 
consecutive  blades  is  made  about 
.5  to  .6  of  the  axial  width.*  In 
Fig.  50  the  pitch  is  shown  by 
the  distance  between  the  points  i 
and  2.  Hence,  when  a  drawing  FlG  49._Diagram  i,lustrating  the  DC- 

of  blades  is  Started   these  points      sign  of  Blades  for  Impulse  Turbines. 

should  first  be  located.     At  any 

point  between  i  and  2  mark  another  point  3  and  through  it  draw 
a  line  33',  making  an  angle  with  the  horizontal  equal  to  the  blade 

angle  y  on  that  side.    Draw 
I  through  2  a  line  perpendic- 

ular to  the  last  line  and 
intersecting  the  center  line 
of  the  blades.  Mark  this 
point  on  the  center  line  5. 
Draw  through  i  a  line  par- 
allel to  3  3'  to  intersect  2  5, 
at  4.  With  5  as  a  center 
draw  an  arc  tangent  to 
i  4,  which  completes  the 
lower  half  of  the  convex 
side  of  the  blade.  With 

the  same  center  the  concave  side  of,  the  next  blade  is  drawn  with 
the  arc  passing  through   2.     The  arrows   in  the   figure   show 

*  The  most  efficient  blade  pitch  appears  to  be  between  the  limits  of  J  inch  and 
i  inch.  Between  these  two  values  the  efficiency  of  blades  made  according  to 
conventional  designs  is  practically  constant.  The  usual  blade  pitches  are  f,  f, 
and  I  inch.  Even  for  very  small  turbines  not  much  less  than  J-inch  pitch  should 
be  used.  Designers  usually  make  the  pitch  and  axial  width  increase  a  little  with 
the  height  of  the  blades. 


FIG.  50.  —  Diagram  Illustrating  the  Method 
f6r  Laying  Out  the  Blades  of  an  Impulse 
Turbine. 


84  THE  STEAM  TURBINE 

plainly  the  center  for  these  arcs.  This  construction  makes  the 
"  perpendicular  "  width  of  the  steam  passage  nearly  constant. 

Blade  Velocity  Losses.  Various  attempts  have  been  made  by 
experimenters  to  determine  the  velocity  losses  in  blades  with 
stationary  apparatus,  usually  by  some  method  of  measuring  the 
reaction  somewhat  in  the  same  way  as  the  coefficients  given  in 
Fig.  28  were  obtained  for  nozzles.  Such  results,  however,  are 
not  satisfactory  for  application  to  designing.  Frictional,  eddy, 
and  impact  losses  in  moving  blades  are  certainly  very  different 
from  what  they  are  in  stationary  blades.  Apparently  there  are 
only  two  ways  to  get  good  data  regarding  these  losses.  Either 
the  velocity  must  be  measured  between  the  blades  of  an  operat- 
ing turbine  with  a  Pitot  tube  or  similar  device,  or  they  must 
be  determined  by  the  "  cut  and  try"  method  that  has  been 
generally  followed  by  designers.  The  latter  method  is  certainly 
expensive  and  a  slow  one  for  obtaining  results.  It  seems,  there- 
fore, that  more  work  should  be  done  along  the  line  of  the  former 
method  by  the  application  of  the  Pitot  tube.  In  the  latest  designs 
of  steam  turbines  there  is  no  difficulty  about  getting  sufficient 
space  for  a  pressure  tube  between  the  blades,  as  the  axial  clear- 
ance in  large  turbines  is  often  as  much  as  ^  inch. 

Fig.  51  shows  values  of  the  velocity  coefficients  to  be  applied 
in  designing  steam  turbine  blades.  Curve  A  is  for  blades  re- 
ceiving steam  from  nozzles.  Curve  B  is  for  stationary  blades, 
or  for  moving  blades  receiving  steam  from  stationary  blades. 
Values  of  both  curves  vary  with  the  relative  velocity  of  the  steam 
in  the  blades.  The  true  velocity  in  the  blades  is  found  by  mul- 
tiplying the  theoretical  relative  velocity  by  the  coefficient  from 
the  curves.*  The  values  given  by  these  curves  may  be  taken 
as  fairly  representative  for  all  the  well-known  commercial  types 

*  Values  given  by  these  curves  agree  well  with  the  determinations  made  by 
Rateau,  Stevens,  and  Hobart  from  the  analysis  of  the  losses  in  operating  turbines. 
Hobart  calculated  that  the  blade  frictional  and  eddy  losses  in  a  275-horsepower 
De  Laval  turbine  are  17  per  cent,  of  the  steam  velocity  which  in  this  case  is  about 
4000  feet  per  second.  He  states  also  that  generally  in  turbines  of  this  type  this 
loss  is  about  15  per  cent,  of  the  relative  velocity  in  the  blades.  It  is  stated  that 
•designers  of  Rateau  turbines  assume  a  blade  velocity  efficiency  of  96  per  cent,  at 
relative  velocities  of  about  600  feet  per  second.  Obviously  near  zero  velocity  the 


STEAM  TURBINE  TYPES  AXD   BLADE   DESIGN 


M 
M 

.so 

.81 
.82 

Ml 

^ 

\ 

s\ 

\ 

\ 

V 

\ 

^s 

N 

\ 

x 

NA 

> 

\ 

N 

X 

N 

SB 

\ 

\ 

N 

X 

v 

^ 

in  which  the  blades  have  smooth  surfaces  and  the  entrance 
edges  are  made  comparatively  sharp  and  at  a  true  angle.  These 
curves  are  intended  to  be  read 
for  only  two  significant  figures.  L00 

The    initial    steam    velocities 
in  turbines  of  the  Parsons  type  |-% 
vary  from  150  to  600  feet  per  |' 
second,  in  Rateau  turbines  from   * 
500  to  1500  feet  per  second,  in  J-88 
Curtis    turbines    from    1500    to 
3000  feet  per  second,  and  in  the 
De  Laval  type  from  2 500  to  4500     ~0      500     1000    1500    2000    2500 

fppf   npr  cpronH  Relative  Velocity  of  Steam 

1Q-  in  Blades      Ft.  per  Sec. 

The  efficiency  of  energy  con- FJG   ^    B,ade    Velodty  Coefficients> 

version  in  the  blades  of  Steam  Curve  A  for  moving  blades  following 
turbines  varies  from  60  to  70  per  nozzles.  Curve  B  for  stationary  blades 

cent.*  for  sizes  from  300  to  3000    or  for  ™°™s  blades  followins  sta' 

tionary  blades. 

kilowattsj  and  is,  roughly  about 

50  per  cent,  for  smaller  sizes  of  impulse  turbines  down  to  about 
10  kilowatts.  Still  smaller  sizes  may  have  efficiencies  less  than 
50  per  cent.,  depending  largely  on  the  type  of  construction.  For 
any  size  of  turbine,  then,  the  difference  between  100  per  cent, 
and  this  efficiency  of  energy  conversion  is  the  loss  due  to  disk  and 
blade  rotation,  leakage,  residual  velocity,  and  radiation. 

In  a  well-designed  turbine  of  say  300  to  500  kilowatts'  capacity > 

loss  is  practically  zero.  Designers  of  Parsons  and  Curtis  turbines  must  use  some- 
what larger  coefficients  (cf.  Curve  B)  for  combinations  of  stationary  and  moving 
blades,  because  stationary  blades  are  not  as  efficient  as  nozzles.  The  data/  for 
these  curves  were  obtained  by  measuring  with  modified  Pitot  tube  apparatus  the 
velocity  of  steam  discharged  from  stationary  blades  of  various  designs.  The 
steam  was  directed  upon  the  blades  from  calibrated  nozzles. 

*  In  stating  this  efficiency  it  is  assumed  that  adequate  provision  is  made  in 
these  turbines  to  prevent  leakage:  in  impulse  turbines,  between  the  diaphragms 
and  the  shaft;  and,  in  reaction  turbines,  over  the  ends  of  the  blades  .through  the 
radial  clearance.  This  leakage  loss  is  as  much  as  10  to  15  per  cent,  in  some  good 
commercial  turbines.  It  should  be  reduced,  however,  to  not  more  than  5  per  cent. 

f  A  well-known  engineer  states  that  the  energy  efficiency  of  the  9000  to  12,000 
kilowatt  turbines  installed  in  New  York  and  Chicago  is  as  high  as  80  per  cent.  On 
a  conservative  basis,  however,  about  70  per  cent,  can  be  assumed  for  5ooo-kilowatt 
sizes  and  75  per  cent,  for  io,ooo-kilowatt  sizes. 


86  THE   STEAM  TURBINE 

the  sum  of  the  losses  due  to  friction,  disk  and  blade  rotation  or 
"  windage/'  leakage,  residual  velocity,  and  radiation  losses  is, 
therefore,  about  40  per  cent.  But  these  are  not  all  actual  losses. 
The  energy  equivalent  of  each  of  these  losses,  except  that  due  to 
radiation,  which  is  very  small,  is  immediately  converted  into 
heat,  and  is  partially  regained  in  reheating  the  steam.  The  sum 
of  these  losses  actually  reheating  the  steam,  expressed  as  a  per- 
centage of  the  total  available  energy,  is  called  the  reheating  factor. 

DESIGN  OF  BLADES  FOR  AN  IMPULSE  TURBINE 

^Blades  are  to  be  designed  for  a  3oo-kilowatt  turbine  to  operate 
}yith  steam  at  50°  F.  superheat,  at  an  initial  pressure  of  165 
pounds  per  square  inch  absolute,  and  exhausting  at  i  pound  per 
square  inch  absolute  (approximately  28  inches  vacuum).  Blade 
speed  V6  is  500  feet  per  second  at  the  rated  speed  3600  r.p.m. 
It  is  assumed  that  the  nozzle  will  be  correctly  designed  for  the 
pressure,  so  that  the  nozzle  velocity  loss  is  3  per  cent.  Governing 
is  to  be  accomplished  by  the  method  of  "  cutting  out  nozzles  " 
in  the  first  stage  (see  page  221).  By  this  method  a  practically 
constant  steam  pressure  is  maintained  in  the  nozzles  of  the  first 
stage  from  light  load  to  overload,  and  the  velocities  in  this  stage 
are  at  all  loads  approximately  those  giving  the  best  blade  effi- 
ciencies. In  the  other  stages,  however,  where  the  number  of 
nozzles  open  is  not  controlled  by  the  governor,  the  velocities  are 
variable.  For  this  reason  a  large  pressure  drop  is  to  be  used 
ior  this  stage,*  and  to  utilize  the  resulting  high  velocity  efficiently 
there  are  to  be  two  velocity  stages  in  this  pressure  stage.  A 
reasonable  value  for  the  first  stage  pressure  is  about  35  pounds 
per  square  inch  absolute.  The  other  stages  are  to  be  designed 
for  highest  efficiency  with  a  single  blade  wheel  in  each  pressure 
stage.  Such  a  design  will  be  a  compound  type  —  the  first  stage 
resembling  the  Curtis,  and  the  other  stages  the  Rateau  turbines. 

The  energy  available  from  adiabatic  expansion  in  the  first  stage 
nozzles  (as  read  from  the  entropy-heat  chart)  from  165  pounds 
per  square  inch  absolute  and  50°  F.  superheat  to  35  pounds 

*  See  footnote  on  page  7 


STEAM  TURBINE  TYPES  AND  BLADE   DESIGN          8/ 

per  square  inch  absolute  is  122  B.T.U.  Disk  and  blade  rotation 
losses,  leakage  between  the  stages  at  the  joint  between  the  shaft 
and  the  diaphragm,  and  residual  velocity  of  the  steam  leaving 
the  blades  amount  to  40  per  cent.;  and  it  is  assumed  that  all 
this  energy  appears  again  as  heat  produced  by  disk  and  blade 
friction,  steam  impact,  eddies,  and  throttling.  There  is  then  40 
per  cent,  of  122  B.T.U.,  or  nearly  49  B.T.U,,  going  to  reheat 
the  steam.  This  reheating  occurs,  of  course,  at  the  pressure 
in  the  first  stage  (35  pounds).  As  the  result  of  reheating,  the 
quality  of  tlje  steam  in  the  first  stage  is  changed  from  .932^0 
.985,  and  the  total  heat  of  the  steam  going  to  the  nozzles  of 
the  next  stage  is  increased  from  1103  to  1152  B.T.U.  Fig.  52 
shows  graphically  this  reheating  effect.  It  serves  also  to  show 


1225  1152  1138   1103  B.T.U.gcale  892 

FIG.  52. — Entropy-Heat  Diagram  for  the  Design  of  an  Impulse  Turbine. 

the  complete  energy  distribution  as  required  for  this  design. 
In  each  stage,  as  in  the  first,  the  reheating  is  assumed  to  be 
40  per  cent. 

.  Since  all  the  stages  after  the  first  are  to  be  of  the  single  wheel 
impulse  type,  it  is  obvious  that  a  large  number  of  stages  will  be 
needed  in  order  to  absorb  the  velocity  of  the  steam  efficiently. 
In  a  stage  of  the  single  wheel  type  the  velocity  of  the  steam  should 


88  THE   STEAM   TURBINE 

not  be  greater  than  twice  the  blade  speed.  Equation  (i8b) 
shows  the  relation  between  the  steam  and  blade  velocities  for 
the  highest  efficiency,  and  this  equation  can  be  used  for  determin- 
ing quite  accurately  the  best  energy  distribution.  Because  a 
designing  coefficient  (C)  must  be  inserted  to  correct  for  the  velocity 
loss  in  the  blades,  this  equation  will  now  be  written 


Now  the  velocity  coefficient  for  single  blade  wheels  is  about 
.95.*  The  angle  /3  is  usually  about  40  degrees.  Blade  speed, 
Vb,  is  500  feet  per  second.  Then 

T7  2    X    500   X   COS  20°  , 

K2  =  -  -  -  =  989  feet  per  second. 

But  from  equation  (2)  we  have  the  relation  that  the  available 
energy  (Ea)  in  terms  of  velocity  is 


V223.7/ 

It  is  shown  then  that  the  required  energy  per  stage  is  between 
19.5  and  20  B.T.U.  per  stage.  The  energy  distribution  with 
reheating  (40  per  cent.)  was  calculated  with  the  help  of  the  chart 
for  19.5  and  for  19.8  B.T.U.  per  stage,  and  it  was  found  possible 
to  get  almost  exactly  equal  energy  distribution  with  1.2  stages 
each  of  19.8  B.T.U.  between  35  pounds  pressure  (quality  .985) 
and  the  exhaust  pressure  i  .o  pound?  This  distribution  is  shown 
in  Fig.  52.  The  quality  of  the  steam  in  each  stage  is  recorded, 
so  that  the  disk  and  blade  friction  can  be  calculated  later  from 
the  formulas  in  Chapter  V. 

*  See  Fig.  51.  To  determine  an  approximate  value  for  this,  coefficient  the 
probable  relative  velocity  must  be  estimated.  If  a  very  large  error  were  made 
in  assuming  this  coefficient  it  would  be  discovered  as  soon  as  the  velocity  diagrams 
are  made,  as  the  relative  velocity  and  the  coefficients  are  then  accurately  deter- 
mined. 


STEAM  TURBINE  TYPES  AND  BLADE  DESIGN 


89 


Velocity  of  the  steam  discharged  .from  the  first  stage  nozzles 
is 

^2  =  -97  X  223-7  Vi22  =?•  2398  feet  per  second, 

and  that  from  each  of  the  other  stages  is  ' 

V2f  =  .97  X  223.7  Vig.8  =  965  feet  per  second. 

The  velocity  coefficients  given  in  Fig.  51  have  been  used  to 
lay  out  the  triangles  in  Figs.  53  and  54.     The  application  can 

be  best  illustrated  by  the 
triangles  in  Fig.  53, show- 
ing the  velocities  of  the 
first  stage. 

For  constructing  the 
triangles  in  Fig.  53,  V2  is 
laid  off  inclined  20  de- 
grees (the  nozzle  angle) 
to  the  horizontal  and 
made  to  scale  2398  feet. 


FIG.  53.   Velocity  Triangles  for  Two  Velocity        FIG.  54.    Velocity  Triangles  for 
Stages  in  One  Pressure  Stage.  a  Simple  Impulse  Wheel. 

To  the  same  scale  the  blade  speed  (F&)  is  laid  off  for  500 
feet,  making  the  relative  velocity  (Vr2)  in  the  first  row  of 
blades  1938  feet  per  second,  and  the  entrance  angle  (B)  of 
these  blades  is  found  to  be  25^  degrees.  The  entrance  and 
discharge  angles  will  be  made  equal,  so  that  the  angle  C  is  also 
25i  degrees,  determining  the  slope  of  the  relative  velocity  (Frs). 
The  velocky  coefficient  taken  from  curve  A  in  Fig.  51  for  a 
relative  velocity  of  1938  feet  is  .88,  so  that  Fr3  =  1938  X  .88  or 
1705  feet.  Vb  is  again  laid  off  in  a  horizontal  direction,  and  the 
absolute  velocity  of  the  steam  discharged  from  the  first  row  of 


90  THE  STEAM  TURBINE 

blades  (F3)  as  read  by  the  scale  is  1270  feet,  and  the  true  dis- 
charge angle  (D)  is  35  degrees.  In  order  that  the  steam  may 
enter  the  stationary  intermediate  blades  without  shock,  the 
entrance  angle  of  these  blades  must  be  also  35  degrees,  and  the 
discharge  angle  (E)  will  be  made  20  degrees,  the  same  as  the  nozzle 
angle.  The  velocity  coefficient  is  now  read  from  curve  B  in 
Fig.  51  for  1270  feet,*  which  is  .87,  and  F4  is  laid  off  for 
1270  X  .87  =  1105  feet.  Completing  the  triangles,  Fr4  is  662, 
and  the  entrance  and  discharge  angles  F  and  G  are  each  35 
degrees.  The  velocity  coefficient  (read  from  curve  B)  is  .93,  so 
that  Fr5  is  615  feet  and  the  final  discharge  velocity  (F5)  is  355 
feet. 

Velocities  and  blade  angles  are  determined  in  the  same  way 
(by  applying  a  velocity  coefficient)  for  the  12  single  wheel  stages 
as  shown  in  Fig.  54. 

Data  and  results  of  these  velocity  triangles  are  tabulated 
below  for  convenient  reference: 

Blade  Angles  and  Velocities  of  First  Stage. 

First  row  (moving):  entrance  and  discharge  angles  25 J  degrees. 

Intermediate  (stationary):  entrance  angle  35  degrees;  discharge 
angle  20  degrees. 

Second  row  (moving):  entrance  and  discharge  angles  35 
degrees. 

Vb   =  500       Fr3  =  1938  X  .88  =  1705    Vr4=  662 

F2   =  2398      F3  =  1270  Frs  =  662  X  .93=615 

Fr2  =  1938      F4  =  1270  X  .87  =  1105    F5   -  355. 

Blade  Angles  and  Velocities  of  Second  to  Thirteenth  Stages. 

Single  row  (moving) :  entrance  and  discharge  angles  39^  degrees. 

F6  =  500  Fr3  =  525  X  .96  =  504. 

F2  =  965  F3  =  339. 

V*  =  525 

A  slightly  higher  efficiency  could  have  been  obtained  if  the  first 
stage  pressure  had  not  been  assumed  but  had  been  determined 

*  In  stationary  blades  the  absolute  and  relative  velocities  are  equal. 


STEAM  TURBINE  TYPES  AND  BLADE  DESIGN          91 

by  a  "cut  and  try"  method  to  get  the  highest  efficiency.  If  the 
energy  for  this  stage  had  been  a  little  less,  the  efficiency  would 
have  been  increased  —  although  an  insignificant  amount.  It  is 
a  good  rule  to  remember  that  with  a  given  blade  speed,  whenever 
the  line  representing  the  residual  velocity  slopes  toward  either 
side  of  the  vertical,  the  minimum  residual  velocity  has  not  been 
found.  A  higher  efficiency  could  have  been  obtained  also  by 
reducing  the  discharge  angle  of  the  intermediate  blades.  This 
angle  is  usually  made  about  the  same  as  the  nozzle  angle  (about 
20  degrees  in  most  types).  If  it  is  made  less  than  20  degrees, 
although  the  apparent  efficiency  will  be  increased,  there  will  be 
probably  a  greater  loss  than  gain  on  account  of  the  steam  spilling 
over  the  blades. 

Stage  Efficiencies.  Nozzle  efficiency  is  assumed  to  be  97  per 
cent.,  on  the  basis  of  the  velocity  developed.  Efficiency  of  the 
energy  conversion  in  the  blades  can  be  calculated  from  the 
results  given  by  the  velocity  triangles  in  Figs.  53  and  54. 

In  the  first  stage  the  velocity  absorbed  in  moving  the  turbine 
is  the  initial  velocity  (F2)  less  the  residual  velocity,  (F5),  and 
the  velocity  losses  in  the  blades  are  (Vn  —  Fr3)  and  (Fr4  —  F^). 
Then  the  energy  absorbed  in  the  first  stage,*  or 

Work   Done  =  F22  -  (Fr22  -  Fr32)  -  (F32  -  F42) 
-  (Fr42-  Fr52)  -  F52. 

Work  Done 

Blade  Efficiency  =  —  — 

Work  Possible 

F,2  -  Fr22  +  Fr,2  -  F32  +  V:~  -  Vr<2  +  VJ  -  V,9 


V? 


Blade  Efficiency  (first  stage) 


(2.SQ8)2  -(iQj8)2  +  (I7Q.08  -  (i27o)2  +  (iio.O2  -  (662)2  4-  (6i.Q2  - 

(2398)2 

Blade  Efficiency  =  75.3  per  cent. 

Nozzle   and   blade   efficiency   of   the    first    stage  is    therefore 
75-3  x  ^97  *  =  74-2  per  cent. 

*  When  writing  efficiency  equations,  it  must  be  remembered  that  efficiencies  are 
proportional  to  the  available  energies  and  to  the  square  of  the  velocities. 


92  THE  STEAM  TURBINE 

Similarly  for  the  second  stage  (also  third  to  thirteenth  stages) 

we  have, 

V  -  —  V  2  —  (V  2  —  V  2} 
Blade  Efficiency  =    -  -  ^  —  ^^  -  rjj- 

Blade  Efficiency  I  <*#  ~  <wY-W  + 


Blade  Efficiency  =  85.3  per  cent. 

Nozzle  and  blade  efficiency  of  the  last  twelve  stages  is  therefore 

85-3  X  V^7  =  84.0  per  cent. 

The  combined  or  "  total  "  nozzle  and  blade  efficiency  of  the 
turbine,  prorated  according  to  stage  energy,  is,  then, 

74.2X122+84-0X19.8X12  =  cent  * 

122  +  (19.8  X  12) 

Besides  the  nozzle  and  blade  losses,  there  are  bearing  losses, 
including  the  friction  of  the  gland  or  stuffing-box  on  the  shaft 
and  the  power  for  the  governor  and  oil  pumps,  amounting  to 
about  2  per  cent  in  a  turbine  of  this  size,  f  The  radiation  loss  is 
about  i  per  cent,  the  loss  due  to  leakage  of  steam  along  the  shaft 
between  the  stages  should  not  be  more  than  7  per  cent.,  and  the 

*  Although  velocity  stages  do  not  give  as  high  net  blade  efficiency,  the  adoption 
of  this  type  for  the  first  stage  makes  it  possible,  because  of  the  large  available  energy 
required  for  this  stage  by  this  method,  to  make  the  turbine  very  economical  at  light 
loads.  By  providing  a  suitable  valve  gear  the  number  of  nozzles  open  in  the  first 
stage  can  be  controlled  by  the  governor.  (See  pages  221-229.) 

f  Bearing  loss  in  turbines  is  usually  very  small.  According  to  Lasche  of  the 
Allgemeine  Electricitat  Gesellschaft,  Berlin,  the  friction  coefficient  (/)  is 

/=  2-  (tXp), 

where  /  is  the  temperature  of  the  bearing  in  degrees  C.  and  p  is  the  pressure  in 
kilograms  per  square  centimeter.  The  rotor  of  a  1000-  kilo  watt  Parsons  turbine 
weighs  about  3000  pounds,  and  the  disks  and  shaft  of  an  impulse  turbine  would 
probably  weigh  less. 

Langen  in  the  Zeitsch.  fur  das  Gesamte  Turbinenwesen  (Oct.  19,  1907)  states 
that  the  bearing  (journal)  friction  of  a  well-designed  Parsons  turbine  is  about 
.2  per  cent.,  and  that  the  total  friction  loss  including  governor  and  oil  pump  rarely 
exceeds  i  per  cent. 

Stodola's  tests  of  a  Zoelly  turbine,  with,  of  course,  a  much  shorter  casing  than 
that  of  a  Parsons  type,  show  the  radiation  loss  from  the  casing  to  be  about  .7  per 
cent. 


STEAM  TURBINE   TYPES  AND   BLADE   DESIGN  93 

actual  net  loss  of  heat  ("  available  ")  due  to  rotation  of  disks  and 
blades  will  be  about  10  per  cent*  The  sum  of  the  bearing, 
radiation,  leakage,  and  rotation  losses  is  then  about  20  per  cent., 
and  the  efficiency  of  the  turbine  as  measured  by  work  done  at 
the  shaft  is  about  81  per  cent,  (from  page  92)  less  20  per  cent., 
or  about  61  per  cent. 

The  theoretical  steam  consumption  (water  rate)  of  a  perfect 
engine  operating  with  steam  at  the  same  initial  pressure,  super- 
heat, and  exhaust  pressure  is  10.24  pounds  per  kilowatt-hour.f 
Since  the  shaft  efficiency  is  61  per  cent.,  the  equivalent  steam 
consumption  per  shaft  kilowatt-hour  developed  in  the  blades  is 
10.24  -f-  .61,  or  1 6.80  pounds.  Generator  efficiency  might  be 
assumed  to  be  about  92  per  cent,  for  a  good  design  suitable  for 
this  high  speed  relatively  to  the  size,  and  the  steam  consumption 
per  kilowatt-hour  "at  the  switchboard"  would  be  about  16.80  -J- 
.92  =  18.26  pounds.f 

The  energy  efficiency,  neglecting  losses,  of  each  stage  with  a 
single  row  .of  blades  can  be  expressed  approximately  by  equation 
(16),  thus, 

Efficiency  =  4  *5°°  (.940  -  522)  =  87  (nearly). 
965     V  9657 

The  nozzles  for  this  turbine  must  be  designed  to  discharge 
at  full  load  300  X  18.26  pounds  or  5478  pounds  per  hour  at 
50°  F.  superheat.  Total  "  throat  area  "  of  the  nozzles  (A0)  can 
be  calculated  by  equation  (8')  for  superheated  steam,  where 

*  The  actual  rotation  loss  for  this  design  can  be  calculated  by  the  formulas 
given  in  the  following  chapter.  But  a  large  part  of  the  "total"  loss  as  calculated 
becomes  again  available  as  the  result  of  reheating.  The  mean  pitch  diameter  of 
the  blades  is 

500  X  60  , 

-  or  2.65  feet. 
3.1416  X  3600 

f  A  kilowatt-hour  is  equivalent  to  2,654,400  foot-pounds  or  3412  B.T.U.  per 
hour  (44,240  foot-pounds  per  minute).  In  this  case  the  total  available  energy 
taken  as  one  expansion  is  (1225  —  892)  333  B.T.U.  per  pound  of  steam,  and  the 
theoretical  steam  consumption  is  3412  -r-  333,  or  10.24  pounds. 

J  Guaranteed  steam  consumption  would  be  about  10  per  cent,  more  than  the 
estimated  water  rate.  It  is  the  usual  practice  of  manufacturers  of  steam  turbines 
and  engines  to  add  a  percentage  of  about  this  value  to  allow  for  possible  defective 
workmanship  in  construction. 


94 


THE    STEAM   TURBINE 


D  =  50  degrees,  F  =*  5478  +  3.600,  .or  1.52  pounds  per  second,, 
and  Pi  =  165  pounds.     Then 

_  60  F(i+. 00065  fl)  _  60  X  1.52  X  1.0325  _.  66         . 
Pi*  (I65)-97 

The  valve  gear  will  be  designed  to  open  8  nozzks  for  the  first 
stage  at  full  load,  with  provision  for  opening  4  more  at  overload, 
so  that  50' per  cent,  overload  can  be  carried  efficiently  by  the 
turbine.  These  nozzles  will  all  be  of  the  same  size.  Each  first 
stage  nozzle  will  have  a  "throat  area""  of  .665  -T-.  8,  or  .083  square 
inch.  It  will  be  assumed  that  the  section  of  the  nozzle  at  the 
throat. is  approximately  square  (with  rounded  corners)  and  that 
its  width  (in  the  radial  direction  with  respect  to  the  blade  disk)  is 
constant  from  throat  to  mouth,  or  is  V.o83,  or  .288  inch. 

A  calculation  should  now  be  made  to  determine  the  height 
of  the  blades  to  give  sufficient  area  for  the  passage  of  the 
steam.  For  this  purpose  the  length  of  the  .nozzles  at  their 
mouths  must  also  be  calculated.  It  is  obvious  that  a  nozzle 


FIG.  55.    Details  of  the  Nozzle  Mouth. 

cannot  be  designed  to  be  cut  off  at  the  end  of  the  expanding 
portion  at  right  angles  to  its  axis;  but  an  extension  or  "tail"  is 
necessary  to  direct  the  steam  upon  the  blades.  To  avoid  spread- 
ing the  jet  -and  making  the  expansion  ratio  uncertain,  this 
"tail"  is  often  made  non-expanding,  so  that  its  wall  is  parallel 
to  the  axis.  The  varying  dimensions  of  the  nozzles  for  this  design 
can  be  determined  then  from  the  expansion  ratio,  which,  accord- 
ing to  the  curve  in  Fig.  21,  is  approximately  1.52  for  the  expan- 
sion in  the  first  stage  nozzles.  Area  at  the  mouth  is  .083  X  1.52 
=.  .1261  square  inch,  but  as  one  dimension  is  constant,  the  longer 


STEAM  TURBINE  TYPES  AND  BLADE   DESIGN  95 

side  of  the  rectangular  mouth  is  .1261  -r-  .288,  or  .438  inch 
(marked  y  in  Fig.  55)  .  By  the  geometry  of  the  figure  it  is  obvious 
that  the  length  z  =  y  -f-  sin  20  degrees,  if  the  nozzle  angle  is 
20  degrees  as  it  is  generally  made.  Then  the  length  of  the 
nozzle  mouth  opposite  the  blades  is  z  =  .438  -r-  .342,  or  1.28 
inches. 

Sufficient  area  must  be  provided  in  the  blades  to  pass  the 
steam  from  the  nozzles.  The  pressure  of  the  steam  in  the 
blades  is  35  pounds  per  square  inch  absolute,  of  which  the  spe- 
cific volume  (dry)  is  11.89  cubic  feet.  The  weight  of  steam 
flowing  per  second  is  1.52  pounds  and  the  volume  (x  =  .985)  pass- 
ing through  the  blades  per  second  is  approximately  1  1  .89  X  .985 
X  1.52  =  17.8  cubic  feet.  This  volume  and  the  velocity  of  the 
steam  determine  the  necessary  height  (h')  of  the  blades.  Net 
area  in  square  feet  between  the  blades  for  passing  the  steam 
from  eight  nozzles  may  be  written  as  (8  X  1.28  X  h'  X  sin  25!°) 
-7-  144  (cf.  page  82).  This  area  multiplied  by  the  relative  veloc- 
ity of  the  steam  in  the  blades  is  another  expression  for  the  vol- 
ume. Now  the  velocity  of  the  steam  in  the  blades,  on  account  of 
the  frictional  losses,  is  variable.  Obviously  the  blade  area  must 
be  made  large  enough  to  pass  the  steam  at  its  lowest  relative 
velocity;  that  is,  when  it  is  discharging  from  the  blades.  The 
height  of  the  first  row  of  blades  in  the  turbine  (Fr3  =  1705)  is 


=  -    -  -         (nearl  } 

8X  i.28Xsm25i°  X  1705 

After  the  angles  and  height  of  -the  blades*  have  been  de- 
termined they  can  be  laid  out  according  to  the  diagram  in  Fig.  50. 

The  effective  blade  area  should  be  calculated  in  the  same  way 
for  every  row  of  blades  in  each  of  the  stages,  as  it  is  very  impor- 
tant that  the  blades  are  provided  with  sufficient  area. 

DESIGN  OF  BLADES  FOR  A  REACTION  TURBINE. 

In  the  design  of  reaction  turbines  of  the  Parsons  type,  many 
rules  or  formulas  must  be  followed  which  have  been  developed  as 

*  The  other  two  rows  of  blades  in  this  same  pressure  stage  are  calculated  from 
equation  above  using  respectively  F4  =  1105  (stationary  blades)  and  Vrs  =  615. 


96  THE    STEAM   TURBINE 

the  result  of  experience  and  have  no  well-defined  scientific  basis. 
Some  of  these  formulas  will  now  be  given  to  determine  the  num- 
ber of  stages,  the  relation  between  the  maximum  blade  height  and 
the  diameter  of  the  rotor,  the  no-load  steam  consumption,  etc. 

The  rotor  of  a  Parsons  turbine,  except  for  marine  services,  is 
made  commonly  in  three  sections  of  different  diameters.  At  the 
high  pressure  end  a  section  of  small  diameter  is  used,  and  the 
intermediate  and  low  pressure  sections  are  made  relatively  larger 
to  allow  for  the  increased  volume  of  the  steam  as  it  expands  in 
the  blades.  In  an  impulse  turbine,  on  the  other  hand,  the  wheels 
in  the  several  stages  are  usually  of  the  same  diameter.  This 
radical  difference  in  the  type  of  construction  results  from  admit- 
ting in  the  reaction  turbine  high  pressure  steam  around  the  whole 
periphery  of  the  rotating  part,  while  in  the  impulse  turbine  the 
admission  steam  is  discharged  through  nozzles  occupying  usually 
only  a  small  part  of  the  periphery. 

The  diameter  of  the  low  pressure  section  of  the  rotor  of  a 
reaction  turbine  is  determined  by  the  permissible  blade  speed 
and  the  rated  speed  of  rotation  (revolutions  per  minute).  With 
a  drum  construction  it  is  not  permissible  to  adopt  peripheral 
speeds  for  the  rotor  higher  than  about  300  feet  per  second.  The 
speed  of  the  rotor  (revolutions  per  minute)  will  depend  on  the 
capacity  of  the  turbine,  or  more  particularly,  if  it  is  to  be  connected 
to  an  electric  generator,  on  the  allowable  speed  of  the  generator. 

A  table  *  on  page  97  gives  the  rated  speeds  of  a  number  of 
different  sizes  of  commercial  turbine-generators  of  the  Parsons 
type,  some  of  which,  it  will  'be  observed,  are  not  for  standard 
frequencies  in  America. 

It  is  much  more  difficult  to  design  a  turbine-generator  with 
sufficient  strength  in  a  rotating  field  or  armature  than  the 
turbine  parts.  The  diameter  of  trie  low-pressure  section  is  gen- 
erally made  \/2  times  that  of  the  intermediate  section,  and  the 
diameter  of  the  intermediate  section  is  \/2  times  that  of  the 
high-pressure  section.  It  follows  then  obviously  that  the  ratio 
of  the  blade  speeds  of  successive  sections  is  also  V2. 

*  Trans.  Inst.  of  Engineers  and  Shipbuilders  (1905-06). 


STEAM  TURBINE   TYPES  AND   BLADE   DESIGN 


97 


The  speed  of  rotation  of  turbines  direct  connected  to  alternat- 
ing-current electric  generators  is  usually  determined  by  the 
frequency.  For  the  usual  frequency  in  America  for  electric 
lighting  (60  cycles)  the  generator  must  be  operated  at  3600, 
1800,  or  900  revolutions  per  minute;  and  for  15  cycles  the  revolu- 
tions cannot,  of  course,  exceed  900  per  minute,  as  a  generator 
cannot  be  built  with  less  than  two  poles. 


Normal  Output  of  Turbine. 

Peripheral  Blade  Speed, 
Feet  per  Second. 

Number  of 
Rows 
of  Moving 
Blades. 

Revolutions 
per 
Minute. 

First 
Expansion 
(Section). 

Last 
Expansion 

(Section).  . 

250  kilowatts  

IOO 
120 

125 
I25 

I25 

I25 

138 
'35 

210 

**s 

260 
250 
300 
300 
280 
330 

72 
60 

77 
80 
72 
84 

75 
70 

3OOO 
3000 
2000 
I800 
1500 
I36o 
I2OO 
750 

500  kilowatts  

750  kilowatts  

i  ooo  kilowatts  

i  coo  kilowatts 

2  coo  kilowatts 

3500  kilowatts 

5000  kilowatts                            .    . 

Usually  in  designs  with  three  different  diameters  of  the  rotor 
(three  sections),  the  number  of  rows  of  blades  or  stages  is 
arranged  so  that  one-quarter  of  the  total  work  is  done  in  the 
high-pressure  section.  The  intermediate  section  takes  also  one- 
quarter  of  the  total  work,  and  the  low-pressure  section  one-half. 

Designers  of  Parsons  turbines  have  long  used  the  following 
formula  to  determine  the  number  of  rows  of  blades : 
Vb2n  =  constant, 

where  Vb  is  the  mean  peripheral  velocity  of  the  blades  of  any 
section  and  n  is  the  corresponding  total  number  of  rows  of 
blades  on  the  rotor  or,  in  other  words,  the  number  of  stages. 
Designers  of  marine  turbines  usually  assume  the  value  of  the 
constant  at  about  1,500,000  to  1,600,000;  but  for  electric 
generator  service,  where  much  higher  peripheral  speeds  are 
allowable,  the  value  of  this  constant  varies  from  2,200,000  to 
2,600,000,  depending  somewhat  on  the  allowable  radial  clear- 
ances. The  lower  value  can  be  used  when  the  machine 
work  is  accurate  and  the  designing  has  been  done  with  great 


98  THE   STEAM  TURBINE 

care  to  eliminate  unequal  expansion  between  the  rotor  and  the 
casing  (see  page  107). 

It  sometimes  happens  when  arranging  the  blading  in  groups, 
that  a  fractional  part  of  a  stage  is  shown  by  the  calculations. 
In  such  a  case  two  groups  may  be  combined  into  one  of  about 
the  average  height,  if  in  this  way  a  whole  number  of  rows  can  be 
secured.  Probably  it  will  then  be  found  that  one  or  two  of  the 
last  rows  of  blades  do  not  give  sufficient  area  for  the  passage  of  the 
steam,  and  this  area  is  then  increased  by  "gauging"  the  blades  in 
both  the  rotor  and  casing.  This  " gauging"  is  done  by  forcing  a 
piece  of  metal  —  preferably  not  much  harder  than  the  metal  of  the 
blades  —  between  the  blades  so  as  to  twist  them  more  nearly  paral- 
lel to  the  axis.  Manufacturers  using  steel  blades  have  usually 
special  keys  made  for  the  purpose  of  twisting  the  blades  by  hand 
both  for  the  purpose  of  " gauging"  and  for  changing  the  blade 
angles  in  order  to  secure  an  accurate  balance  between  the  end 
thrust  of  the  balance  pistons  (see  page  156)  and  that  of  the  blades. 
It  is  stated  on  very  good  authority  that  this  twisting  of  the  blades 
and  changing  the  angles  with  respect  to  the  steam  flow  as  much  as 
5  degrees  does  not  appreciably  alter  the  economy  of  the  turbine. 

An  example  illustrating  the  design  of  a  commercial  type  of 
reaction  turbine  will  now  be  discussed. 

The  difficult  part  and  that  requiring  the  best  judgment  in  the 
designing  of  a  reaction  type  of  steam  turbine  is  in  determining 
as  accurately  as  possible  the  volume  of  steam  that  will  pass 
through  the  blades  for  its  full  capacity;  that  is,  when  all  the 
valves  controlling  the  admission  of  steam  are  wide  open;  or  in 
other  words  when  there  is  no  throttling  of  the  steam  pressure. 
It  is  for  this  flow  that  all  the  blades  must  be  proportioned  for 
their  best  efficiency.  It  is  presumed  that  for  both  lighter  and 
heavier  loads  the  efficiency  and  the  steam  consumption  will  not 
be  so  good.  In  order  to  determine  the  volume  of  steam  flowing 
at  this  condition  obviously  the  actual  number  of  pounds  of 
steam  to  be  used  by  the  turbine  must  first  be  known. 


STEAM  TURBINE   TYPES  AND   BLADE   DESIGN 


99 


As  the  result  of  a  great  deal  of  study  the  curve  shown  in 
Fig.  55a  has  been  developed  from'  data  collected  in  a  large  part 
by  Martin  *  as  applying  to  a  large  variety  of  turbines,  but  par- 
ticularly to  the  reaction  type.  This  curve  shows  by  its  ordinates 
the  so-called  "  efficiency  ratio  "  which  is  the  ratio  of  the  theo- 
retical steam  consumption  (see  page  93)  to  the  actual  steam 


50,000 


100,000      150,000 

Coefficient  "C" 


200,000 


FIG.  55a.    Efficiency  Ratios  for  Reaction  Turbines. 

consumption  (water  rate)  per  kilowatt-hour.  The  values  on  this 
curve  from  110,000  to  120,000  are  for  reaction  turbines  either 
above  5000  kilowatts'  capacity  or  else  for  sizes  between  1500  to 
5000  kilowatts,  having  clearances  at  the  tips  of  the  blades  too 
small  for  standard  practice.  Many  of  the  latter  turbines  did,  in 
fact,  strip  their  blades  a  short  time  after  being  put  into  service. 
Values  less  than  100,000  are  for  sizes  smaller  than  1500  kilo- 
watts. Other  things  being  equal  the  smaller  size  turbine  should 
have  a  lower  value  of  the  coefficient.  Thus  for  a  turbine  of 
1000  kilowatts'  capacity  the  proper  coefficient  should  be  between 
80,000  and  90,000  and  for  a  2000  kilowatt  size  the  coefficient 
should  be  about  100,000.  In  fact  the  latter  value  is  generally 
used  by  careful  designers  of  Parsons  types  for  all  sizes  from 

*  Design  and  Construction  of  Steam  Turbines,  1913. 


100  THE  STEAM  TURBINE 

1500  to  5000  kilowatts,  having  reasonably  large  clearances  at 
the  tips  of  the  blades. 

Practical  Example.  A  reaction  turbine  with  a  drum  rotor  of 
three  sections  is  to  be  designed  to  give  a  rated  output  of  2000 
kilowatts,  operating  at  1500  r.p.m.  When  supplied  with  steam 
at  165  pounds  absolute  pressure,  100°  F.  superheat,  and  i  pound 
absolute  exhaust  pressure  (about  28  inches  vacuum),  the  turbine 
shall  carry  15  per  cent,  overload  before  the  by-pass  or  overload 
valve  (see  page  170)  opens.  , 

For  this  design  (2000  kilowatts),  therefore,  the  value  of  the 
coefficient  should  be  100,000.  Ordinates  of  the  curve  in 
Fig.  55a  show  that  the  corresponding  efficiency  ratio  is  about 
..675. 

U  The  available  energy  from  the  entropy- total  heat  chart  from 
the  initial  conditions  of  165  pounds  per  square  inch  absolute 
pressure  and  100°  F.  superheat  to  the  final  pressure  of  one  pound 
absolute  is  1252  —  908  or  344  per  pound  of  steam.  Dividing 
the  B.T.U.  equivalent  of  a  kilowatt-hour,  which  is  3412,  by  the 
available  energy  per  pound  of  steam  (344)  we  obtain  a  theoreti- 
cal steam  consumption  of  9.92  pounds.  The  theoretical  steam 
consumption  divided  by  the  efficiency  ratio  gives  the  actual 
steam  consumption  of  the  turbine  per  kilowatt-hour  as  meas- 
ured at  the  turbine  shaft  or  9.92  -r-  .675  is  14.70  pounds.  If  we 
assume  5  per  cent,  loss  for  generator  and  connections  to  the 
switchboard  then  the  steam  consumption  per  kilowatt  "  at  the 
switchboard  "  is  14.69  -r-  .95  or  15.46  pounds  of  steam  when  dry 
saturated  (no  superheat). 

The  steam  consumption  of  reaction  turbines  is  reduced  at 
least  10  per  cent,  if  the  steam  is  superheated,  as  in  this  case 
100°  F.  (see  page  281),  so  that  the  actual  number  of  pounds  of 
steam  to  be  passed  through  the  turbine  for  the  conditions  stated 
for  this  design  is  15.46  X  .90  or  13.91  pounds  per  hour  per  kilo- 
watt "  at  the  switchboard." 

The  turbine  must  be  designed  for  a  total  steam  consumption  of 
13.91  X  2000  X  1.15  =  31,993  pounds  per  hour  or  8.88  pounds 
per  second  at  maximum  output,  when  the  admission  valve  will 


STEAM   TURBINE   TYPES  AND   BLADE   DESIGN  iooa 

be  wide  open  so  that  there  is  no  throttling.  Then  the  steam 
entering  the  first  row  of  blades  will  be  at  165  pounds  absolute 
pressure,  of  which  the  volume  at  100°  F.  superheat  is  3.21* 
cubic  feet  per  pound.  The  volume  of  steam  admitted  to  the 
turbine  per  second  is  3.21  X  8.88  =  28.50  cubic  feet,  and  just 
as  in  the  design  of  impulse  turbines,  the  blades  must  be  designed 
for  the  passage  of  this  amount  of  steam. 

The  blades  are  designed  by  determining  the  entrance  and  dis- 
charge angles  by  velocity  triangles  like  those  in  Fig.  4.7  after 
the  available  energy  for  each  stage  has  been  calculated.  Some 
designers  make  their  calculations  for  the  rated  full  load  condi- 
tions'and  not  for  the  maximum  output  obtained  just  before  the 
stage  valve  opens.  The  difference  between  the  two  methods  is 
that  until  the  maximum  output  is  reached,  without  opening  the 
stage  valve,  there  is  obviously  some  throttling  in  the  admission 
valve, f  and  when  designing  for  full  load  conditions  this  throttling 
must  be  allowed  for.  For  this  reason  it  is  preferable  to  design  for 
maximum  output  when  the  admission  valve  must  be  wide  open.f 
The  available  energy  is  then  calculated  by  steps  from  the  rated 

*  Marks  and  Davis'  Steam  Tables  and  Diagrams.  In  these  tables  the  specific 
volumes  have  been  calculated  by  Knoblauch's  equation,  which  gives  considerably 
larger  values  than  equation  (9).  The  results  of  different  investigations  do  not 
give  any  sort  of  agreement,  the  rate  of  increase  of  volume  with  superheating  vary- 
ing as  much  as  100  per  cent.  It  is  usually  stated  that  the  specific  volume  of  super- 
heated steam  is  15  per  cent,  larger  for  100°  F.  of  superheat  than  that  of  dry  saturated 
steam.  According  to  Knoblauch's  equation  used  by  Peabody,  this  percentage  is 
about  17,  and  according  to  equation  (9)  it  is  about  13. 

f  There  is  some  throttling  even  in  the  "pulsating"  valves  used  in  nearly  all 
types  of  Parsons  turbines. 

J  The  steam  consumption  at  fuil  and  fractional  loads  can  be  estimated  by 
drawing  a  "  Willans"  line  of  total  steam  per  hour  (page  124).  Unless  the  design, 
of  a  steam  turbine  is  radically  wrong,  usually  because  of  insufficient  area  of  the 
steam  passages,  which  is  called  "  choking"  the  steam,  it  has  been  shown  by  expe- 
rience that  the  points  representing  total  steam  per  hour  plotted  against  fractional 
loads  will  be  on  a  straight  line  from  no  load  to  the  maximum  output  (without  a, 
stage  valve.  At  no  load  a  Parsons  turbine  usually  takes  one-eighth  of  the  total 
quantity  required  at  the  normal  maximum  output.  By  plotting  these  two  points 
(no  load  and  maximum  output)  and  joining  them  with  a  straight  line,  the  total 
steam  consumption  at  all  other  loads  can  be  read  and  the  steam  per  kilowatt-hour 
or  per  horsepower-hour  can  be  calculated  with  considerable  accuracy. 


loob  THE  STEAM   TURBINE 

admission  to  the  exhaust  pressure.  This  available  energy  might 
be  determined  for  every  stage  as  it  is  done  for  designing  impulse 
turbines,  but  this  is  unnecessarily  laborious,  as  the  pressure  ctrop 
is  so  small.  Approximately  the  same  result  is  obtained  by  cal- 
culating assumed  expansions  in  stages  of  10  B.T.U.  with  the 
same  reheating  factors  as  would  be  used  for  the  same  size  of 
impulse  turbine.  For  a  2000  to  3000  kilowatt  size  the  reheat- 
ing factor  should  be  not  much  more  than  30  per  cent.  Assuming 
this  value,  the  total  available  energy  as  read  from  the  entropy- 
heat  chart  with  reheating  for  every  10  B.T.U.  from  165  pounds 
absolute  and  100°  F.  superheat  to  i  pound  absolute  exhaust  is 
260  B.T.U.*  Without  considering  reheating  it  would  have  been 
343  B.T.U.;  but  with  30  per  cent  reheating  it  is  only  240  B.T.U. 
The  quality  of  the  steam  in  the  last  stage  after  reheating  "  by 
steps  "  is  .886. 

For  this  design  it  will  be  assumed  that 

Vb*n  =  2,560,000. 

It  has  already  been  stated  that  the  diameters  of  the  sections  of 
the  rotor  increase  as  A/2 ;  and  as  the  blade  speeds  must  increase  in 
the  same  proportion  as  the  diameters,  the  following  speeds  of  the 
blades  will  be  assumed,  which  are  not  at  variance  with  good  prac- 
tice: 

Vb  of  first  section  of  rotor  =  140  feet  per  second. 

Vb  of  second  section  of  rotor  =  200  feet  per  second. 

Vb  of  third  section  of  rotor  =  280  feet  per  second. 

The  value  of  peripheral  speed,  140  feet  per  second  for  the  first 
section  of  the  rotor,  corresponds  at  the  speed  of  rotation  required 
(1500  revolutions  per  minute)  to  a  diameter  of 

=  1.78  feet  or  21.36  inches. 

3.1416  X  1500 

*  This  available  energy  should  be  read  in  the  same  way  as  for  the  design  of  the 
impulse  turbine  illustrated  in  Fig.  52;  meaning,  that  the  energy  should  be  obtained 
by  subtracting  from  the  total  heat  at  the  initial  condition  of  pressure  and  super- 
heat, the  total  heat  at  the  final  pressure,  without  the  last  reheating.  There  are  some 


STEAM   TURBINE   TYPES   AND   BLADE    DESIGN  IOOC 

It  is  stated  by  Martin  that  English  designers  of  reaction  turbines 
for  land  service  between  1000  and  6000  kilowatts'  capacity  deter- 
mine the  diameter  of  the  first  section  of  the  drum  of  the  rotor  di 
by  the  following  empirical  formula  based  on  experience: 

^  _  410,000  wv0^ 
r.p.m. 

where  w  is  the  weight  of  steam  flowing  through  the  turbine  in 
pounds  per  second  at  maximum  output  without  the  stage  valve 
being  open,  and  v0  is  the  specific  volume  of  the  steam  at  the  con- 
dition it  enters  the  turbine  in  cubic  feet  per  pound.  In  this  case 
the  diameter  of  the  first  section  of  the  rotor  as  calculated  by  this 
formula  would  be 

di,  =  4.0.00°  X  8.88  X  3.21  = 

I5OO 

and  di  is  ^7790  or  19.8  inches,  which  agrees  well  with  the  value 
calculated  (21.36  inches)  from  the  assumption  of  a  satisfactory 
peripheral  speed.  Actually  it  is  better  practice  and  certainly 
more  rational  to  assume  a  safe  peripheral  speed  than  to  deter- 
mine the  diameter  by  formulas  having  important  empirical 
coefficients  which  must  vary  necessarily  considerably  with  the 
type  of  the  design.  Allowable  peripheral  or  blade  speeds  are 
always  about  the  same  for  a  given  speed  and  type  of  construc- 
tion. Limits  as  regards  peripheral  speeds  can  be  as  accurately 
determined  as  any  other  problem  in  the  designing  of  machines. 
A  similarly  empirical  formula  is  sometimes  used  by  designers 
of  reaction  turbines  to  determine  the  least  permissible  diameter 
of  the  low-pressure  section  in  the  last  stage,  with  the  object  of 
reducing  to  a  minimum  the  losses  due  to  excessive  residual 
velocity  of  the  steam  as  it  discharges  into  the  exhaust  pipe.  If 
the  diameter  of  the  rotor  measured  to  the  middle  of  the  blades 
in  the  last  stage  is  dz  then  for  this  reason  d22  should  be  not  less 

designers  of  impulse  turbines,  however,  who  use  the  calculated  net  available  energy 
after  reheating  in  each  stage;  but  the  problem  then  becomes  very  complicated, 
as  most  of  the  reheating  takes  place  after  the  steam  is  discharged  from  the  nozzles 
or  stationary  blades. 


lood  THE    STEAM   TURBINE 

than  .57  X  the  output  when  there  is  no  throttling  in  the  main 
inlet  valve.  In  this  case  then 

d*2  >  -57  X  2000  X  1.15  (see  page  100), 

or  dzmust  be  at  least  33.76  inches,  which  corresponds  to  a  periph- 
end*  speed  of  33-?6  X  3.1416^  X  1500  >  m  ^  ^  ^  ^^ 

The  value  selected  for  this  stage  (280  feet  per  second)  from  the 
viewpoint  of  permissible  stresses  is  well  above  this  minimum 
limit.  Blade  speeds  as  high  as  400  feet  per  second  are  now  used 
in  some  American  designs  of  steam  reaction  turbines. 

Having  determined  that  the  conventional  blade  speeds  are 
very  satisfactory  for  this  design  the  required  number  of  reaction 
stages  will  be  calculated  in  the  usual  manner  as  follows: 

If  the  blade  speed  of  the  whole  turbine  had  a  constant  value  of 
140  feet  per  second,  then 

(i4o)2  X  n  =  2,560,000;  n  =  128  (nearly). 

As,  however,  only  one-fourth  of  the  work  is  to  be  done  by  the 
first  section  *  operating  at  this  blade  speed,  the  number  of  stages 

in  the  first  section  is  -  -  =  32.     The  value  of  n  for  the  second 
4 

section  is  64;  and  as  one-fourth  of  the  work  is  done  also  in  this 
section,  the  number  of  stages  is  16.  For  the  third  section  n  is 
32,  and  since  one-half  of  the  work  is  done  in  this  section,  the 
number  of  stages  is  16. 

Each  section  of  the  rotor  is  commonly  divided  into  two  or 
four  groups  or  "  expansions." 

Reaction  turbines  are  usually  designed  for  equal  work  (energy) 
per  stage  for  a  given  section  of  the  rotor.  In  the  high-pressure 
or  first  section,  one-quarter  of  the  work  is  done,  and  the  avail- 
able energy  for  each  of  its  thirty-two  stages  is  260  B.T.U.  -r- 
(4  X  32)  =  2.03  B.T.U.  Similarly  the  available  energy  for 
each  stage  of  the  intermediate  section  is  260  B.T.U.  -r-  (4  X  16) 
=  4.06  B.T.U.;  and  for  each  stage  of  the  low-pressure  section 
is  260  B.T.U.  -*-  (2  X  16)  =  8.13  B.T.U.  It  may  be  assumed 

*  See  page  97. 


STEAM   TURBINE  TYPES  AND  BLADE   DESIGN  IOI 

that  about  one-half  of  the  available  energy  in  each  stage  produces 
velocity  in  the  stationary  blades'  and  the  other  half  in  the  moving 
blades.  The  theoretical  angles  are  determined  from  velocity  tri- 
angles, applying  the  coefficients  from  curve  B  in  Fig.  51,  by  the 
usual  methods  as  explained  for  impulse  turbines.  The  discharge 
angles  for  all  the  stages  except  the  last  groups  in  the  low-pressure 
section  will  be  assumed  to  be  20  degrees.  The  angles  for  the  last 
stages  will  be  made  45  degrees.  It  is  obvious,  of  course,  that 
the  discharge  angle  is  always  the  same  as  the  "  absolute  "  angle 
at  which  the  steam  enters  the  succeeding  row  of  blades.  In  this 
design  no  allowances  are  made  for  probable  "  gauging  "  of  the 
blades  to  adjust  the  thrust  on  the  rotor  or  for  other  reasons.* 
The  velocity  of  the  steam  leaving  the  first  row  of  stationary 
blades  in  the  high-pressure  section  is  about  225  f  feet  per  second. 
A  net  area  of  28.50  cubic  feet  -f-  225,  or  .127  square  feet,  or  18.2 
square  inches,  is  required  to  pass  the  steam.  As  the  discharge 
angles  of  the  blades  in  the  high-pressure  and  intermediate  sec- 
tions are  to  be  made  20  degrees,  that  value  will  be  taken  for  this 
design,  and  the  actual  area  of  the  blade  ring  will  be  approximately 
1 8. 2  -f-  .342, t  or  53.3  square  inches. 

The  blade  speed  of  the  high-pressure  rotor  is  140  feet  per 
second,  §  so  that  the  mean  diameter  of  the  blade  ring  is 

*  It  has  been  stated  that  some  makers  of  marine  turbines  who  have  not  had 
much  experience  in  building  them  will  often  design  turbines  to  give  considerably 
larger  output  than  is  intended  for  the  service  and  then  reduce  the  output  to  the 
required  rating  by  "  gauging"  the  blades. 

t  In  a  reaction  turbine  the  maximum  velocity  in  each  stage  is  attained  when  the 
steam  is  discharged  from  the  stationary  blades.  Although  there  is  expansion  also 
in  the  moving  blades,  more  velocity  is  absorbed  in  them  than  is  produced,  and  the 
velocity  of  the  steam  discharged  from  the  moving  blades  is  considerably  less  than 
225  feet  per  second. 

J  The  total  area  of  the  annulus  for  blades  with  discharge  angles  of  20  degrees  is 
the  net  required  area  divided  by  sin  20  degrees  (see  Fig.  49).  Practical  designers 
often  call  the  sin  of  20  degrees  one-third  and  make  the  area  of  the  annulus  three 
times  the  net  required  area. 

§  Manufacturers  generally  appreciate  the  gain  from  operating  at  high  peripheral 
speeds  of  the  rotor.  To-day  efforts  are  directed  generally  by  all  makers  of  direct- 
connected  turbine-generators  to  improve  the  mechanical  construction  of  the  gen- 
erator to  run  at  higher  speeds. 


102  THE   STEAM   TURBINE 

— 14°  X  =  1.78  feet,  or  21.4  inches,  and  the  height  of  the 

1500  X  3-I4l6 

first  row  of  blades  on  the  rotor  is  approximately  53.3  square  inches 
-T-  21.4  X  3.1416  =  .80  or  nearly^f  (see  table,  page  103)  inch. 

With  full  rated  pressure  in  the  admission  chamber  about  7  per 
cent,  of  the  total  steam  leaks  through  the  " dummies"  or  balance 
pistons  at  the  high-pressure  end  of  the  turbine.  This  leakage 
as  well  as  that  around  the  tips  of  the  blades  through  the  radial 
clearance  is  not  considered  here  in  the  calculations.  It  is  prob- 
able, however,  that  the  amount  of  this  leakage  is  quite  sufficient 
to  allow  for  the  thickness  of  the  blades  on  the  discharge  side.* 
The  volume  of  the  exhaust  steam  (i  pound  per  square  inch  abso- 
lute pressure  and  .886  quality)  is  297  cubic  feet  per  pound. 
Initially  the  volume  was  3.215  cubic  feet  per  pound,  so  that  the 
volume  in  the  last  row  of  blades  is  92.5  times  that  at  admission. 
Since  one-fourth  of  the  work  is  done  in  the  blades  of  the  first 
section,  one-fourth  of  the  total  expansion  occurs  in  them,  or  the 
volume  entering  the  second  section  is  ^92.5^  or  3.10  tunes  the 
original  volume.  Since  the  mean  diameter  is  to  be  made  A/2 
times  that  at  the  high-pressure  end  and  the  steam  velocity  is  to 
be  also  A/2  tunes  as  great  so  as  to  correspond  with  the  increase 
in  blade  speed  (which  is  \/2  times  that  in  the  first  section.  See 
page  100),  the  height  of  the  blades  in  the  first  row  of  the  inter- 
mediate section  will  be  3-10  =  1.55  times  that  of  the  first 

V  2   X   V  2 

row  in  the  high-pressure  section.  Similarly  the  blade  height  for 
the  first  row  of  the  low-pressure  end  will  be  1.55  times  that  of  the 
first  row  of  the  intermediate  section.  Each  of  these  sections  will 
be  divided  into  four  groups  or  "  expansions."  Since  the  volume 
is  increased  four  times  for  each  section,  the  blade  height  of  each 

*  Thomas  uses  a  coefficient  of  1.5  to  increase  the  area  of  the  blades  to  allow  for 
the  thickness  at  the  discharge  side.  If  the  blades  are  made  thin  at  their  edges,  as 
in  good  designing,  it  is  not  customary  to  use  a  coefficient  "  for  the  thickness  of  the 
blades." 

t  Let  vr  =  volume  at  end  of  third  section, 

vi  =  volume  at  beginning  of  first  section, 
x  =  number  of  expansions, 
then         vr  =  vf,  and  v\  =  *ff. 


STEAM  TURBINE  TYPES  AND   BLADE   DESIGN 


103 


of  the  high-pressure  and  intermediate  groups  will  be  ^3.10,  or 
1.33  times  as  large  as  in  the  preceding  one. 

The  calculated  blade  heights  for  each  of  the  four  groups  of  the 
high-pressure  and  intermediate  sections  are  given  in  the  follow- 
ing table: 


Group  Number. 

i 

2 

3 

4 

Blade  height, 
Blade  height, 

high-pressure  section 
intermediate  section 

f 

\t 

$ 

lit 
»« 

Blade  heights  are  adjusted  to  sixteenths,  although  in  practice 
the  nearest  eighth  is  commonly  used. 

Because  of  the  long  blades  in  the  low-pressure  section  they  will 
be  made  in  eight  groups.  The  height  of  the  first  group  will  be 
1.55  times  the  height  of  the  first  group  of  the  intermediate  section. 

The  volume  entering  the  third  section  is  ^92.5  X  ^92.5  = 
9.61  times  the  original  volume,  and  blade  height  in  first  row  of 
third  section  is  1.55  X  1.55  X  height  of  first  row  =  (i.ss)2  X  .80 
=  1.9  inches,  or  approximately  iH  inches. 

Each  blade  in  third  section  is  ^9.61  =  1.33  times  height  of 
preceding  one,  or  the  height  of  second  row  is  1.33  X  i|f  or  2.57 
(approximately  2T\  inches) .  The  results  are  tabulated  as  follows : 

Group  Number. 


i 

2 

3 

4 

5 

6 

7 

8 

Blade  height  (inches) 

itf 

*& 

3* 

4& 

6& 

8| 

ioi 

Ml 

Martin  states  that  at  the  high-pressure  end  (in  turbines  for 
stationary  service)  it  is  desirable  to  limit  the  blade  height  to 
not  less  than  one- twenty-fifth  (&)  of  the  drum  diameter.  If 
the  blades  are  shorter  than  this  the  loss  by  leakage  around  the 
tips  may  become  excessive.  In  marine  turbines  the  high-pressure 
bla.des  in  the  first  section  are  only  7^  of  the  drum  diameter. 
On  this  basis  the  blade  heights  might  be  slightly  increased. 


104 


THE  STEAM  TURBINE 


At  the  low  pressure  end  of  the  turbine  the  length  of  the  blades 
would  be  considered  excessive  in  practice.  It  is  a  rule  generally 
followed  by  designers  of  reaction  turbines  not  to  make  the 
greatest  blade  height  more  than  one-sixth  the  mean  diameter  of 
the  blades  for  the  section  considered.  The  mean  diameter  of 
the  low  pressure  section  is  21.4  X  \/2  X  \/2  =  42.8  inches. 
In  this  case  the  maximum  height  would  be,  therefore,  about  7.1 
inches.  In  order  to  reduce  the  length  of  the  blades  so  that 


FIG.  56.     Details  of  the  Design  of  Reaction  Blades. 

practical  requirements  shall  not  be  exceeded,  the  discharge  angle 
of  the  blades  must  be  made  greater  than  20  degrees.  Such  blades 
with  enlarged  "exit"  angles  are  called  wing  blades.  The  tangent 
to  the  curve  at  the  back  of  the  blade  on  the  entrance  side  becomes 
about  90  degrees,  and  at  the  discharge  side  45  degrees  instead 
of  20  degrees.  As  the  result  of  this  change,  the  net  area  for  the 
passage  of  the  steam  is  .71  *  (sin  45  degrees)  instead  of  the 
standard  "£"f  °f  tne  annulus  without  blades.  Wing  blades 
7  inches  long  can  be  used  to  replace  satisfactorily  the  blades  in 
the  5th  group;  but  as  those  of  the  j5th,  7th,  and  8th  groups  must 

*  In  the  turbines  of  the  steamer  Mauretania,  wing  blades  giving  a  passageway 
of  .86  of  the  annulus  were  used,  but  such  a  large  degree  of  "-winging  "  is  not 
adopted  in  steam  turbines  for  electric  generators. 

t  The  sin  of  20  degrees  is  .34,  but  practical  designers  take.it  often  for  convenience 
in  calculating  as  J. 


STEAM  TURBINE  TYPES  AND   BLADE   DESIGN 


105 


be  made  of  the  same  length,  the^e  blades  will  be  shorter  than  they 
should  be.  This  constriction  of  the  steam  passage,  however,  can- 
not well  be  avoided  without  making  the  rotor  in  four  diameters. 
Fig.  56  shows  how  the  blades  of  reaction  turbines  are  laid  out. 
As  explanatory  of  this  figure  a  table  is  given  below  showing  the 
corresponding  dimensions  used  by  one  manufacturer.*  In  the 
table  data  for  five  standard  blades  are  given  for  varying  discharge 
angles  (0)  from  20  degrees  to  35  degrees  and  blade  widths  (w) 
of  .25,  .375,  and  .50  inch.  All  the  linear  dimensions  are  given 
in  inches.  ft  is  the  entrance  angle  of  the  blades. 

Blade  Number. 


< 

2 

4 

5 

0 

20 

2O 

20                              30 

5 

a 

10° 

9°  30' 

1  4°  30' 

15°  45' 

1  8®  40' 

w 

0.25 

0-375 

0.50 

0.50                 0.50 

ft 

67°  30'    • 

67°  30' 

.  67°  30' 

60° 

60° 

R 

0.485 

o-555 

0.794 

0.804 

0.810 

A 

°-°35 

0.045 

0.068 

0.050 

0.040 

b 

0.020 

O.O2O 

O.O2O 

0.020 

0.020 

R, 

0.172 

0.260 

0.342 

0.3*3                °-3°4 

C 

0.049 

O.O4O 

O.I  IO 

0.147 

0.218 

R, 

0.070 

0.109 

0.164 

0.210 

0.212 

k 

0.008 

0.010 

0.015 

O.O4O 

0.056 

m 

0.123 

0.185 

0.288 

0.280 

0.280 

n 

I 

0.185 
o.i  66 

0.280 

0.223 

0.383 

0.282 

0.332 
0.156 

0-33° 
o  134 

H 

0.478 

0.552 

0.770 

0.770                       0.770 

! 

Another  table  is  given  here  showing  the  principal  dimensions 
of  a  4oo-kilowatt  reaction  turbine  with  3,  4,  and  5  groups  per 
section.  The  blade  numbers  in  this  table  refer  to  the  corre- 
sponding numbers  in  the  table  above.  This  table  is  partic- 
ularly useful  for  showing  values  assumed  by  designers  for  the 
blade  pitch.  It  is  not  considered  practicable  in  this  type  of  blade 
construction  to  use  a  smaller  pitch  than  .177  inch  when  a  calking 
tool  must  be  inserted  between  the  blades.  Manufacturers  have 
usually  curve  sheets  of  empirical  data  from  which  the  pitch  and 
other  standard  dimensions  are  obtained. 


*  The  Engineer,  Dec.  16,  1907. 


io6 


THE  STEAM  TURBINE 


FIRST    SECTION. 


Number 
of 
Group. 

Diameter  of 
Section  in 
Feet. 

Dis- 
charge 
Angle. 

Blade 
Height  in 
Inches. 

Blade 
Number 

Volume 
Cubic  Feet 
per  Pound. 

vb+  i\ 

Blade 
Pitch. 

Number 
of 
Blades. 

I 

0.84 

20 

0.6875 

, 

4.08 

.62 

0.177 

179 

5-6 

0.25 

127 

2 

0.84 

20 

1.  00 

I 

5-6 

.62 

0.1875 

169 

7.38 

0.2475 

128 

3 

0.84 

2O 

'•*5 

I 

7.38 

.62 

0.172 

184 

8.92 

0.2175 

146 

SECOND    SECTION. 


I 

1.187 

20 

0.6875 

2 

8.92 

.62 

O.2O 

180 

10.63 

0.31 

116 

2 

1.187 

20 

0-9375 

2 

10.63 

.62 

0.215 

207 

13.8 

0.3075 

144 

3 

1.187 

20 

1.2 

2 

13.8 

.62 

0.215 

207 

4 

1.187 

20 

1-75 

2 

18.8 
18.8 

.62 

0.323 
0.208 

158 
214 

26.6 

0.326 

i37 

THIRD    SECTION. 


I 

1.88 

20 

0-9375 

2 

26.6 

.62 

0.208 

340 

35-8 

0.307 

230 

2 

1.88 

20 

L3I25 

2 

35-8 

.62 

0.208 

34° 

52.8 

0-339 

210 

3 

1.88 

20 

2.00 

2 

y-\ 

.62 

0.198 

358 

83.8 

0-355 

200 

4 

1.88 

30 

2.75 

4 

83.8 

.69 

0.25 

284 

161 

•55 

0-53 

134 

5 

1.88 

30 

4.5 

4 

"161 

.70 

0.308 

230 

311 

.46 

0.54 

J31 

Radial  Leakage.  As  the  volume  of  the  steam  increases,  the 
area  of  the  annulus  of  each  ring  of  blades  must,  of  course,  increase 
proportionally.  This  increased  area  would  be  obtained  most 
easily,  as  with  impulse  turbines,  by  increasing  the  blade  heights 
in  each  ring.  This  method,  however,  would  make  it  necessary 
to  carry  as  stock  in  the  store-room  a  great  number  of  blades  of 
different  sizes.  In  order  to  reduce  the  stock  of  blades  and  to 
reduce  the  cost  of  machining  the  rotor  and  casing,  it  is  custom- 
ary to  make  a  group  of  several  rows  of  blades  of  the  same  height, 


STEAM  TURBINE  TYPES  AND   BLADE   DESIGN          1 07 

and  the  required  increase  in  area  through  each  ring  of  blades  is 
obtained  by  decreasing  the  number  of  blades  in  each  succeeding 
stage.  The  two  values  of  volume,  pitch,  and  number  of  blades 
given  for  each  group  in  the  preceding  table  are  for  the  rows  at  the 
beginning  and  at  the  end  of  the  group. 

In  the  discussion  of  the  design  of  these  reaction  turbines  it  has 
been  assumed  that  each  section  of  the  rotor  is  made  of  the  same 
diameter  from  the  first  to  the  last  group.  For  theoretical  con- 
siderations this  assumption  is  permissible,  but  actually  for  each 
blade  group  the  diameters  of  both  the  rotor  and  casing  are 
changed  so  that  approximately  half  the  increase  in  blade  height 
is  cut  out  of  the  casing  and  the  other  half  is  taken  from  the  rotor. 
It  is  usually  stated  that  this  is  done  merely  for  mechanical  reasons, 
but  this  method  has  advantages  also  in  order  to  secure  the  best 
steam  flow.  It  is  well  known  that  steam  tends  to  fill  completely 
the  passage  through  which  it  flows  and  at  the  same  time  expand 
at  right  angles  to  its  axis  of  flow.  Now  if  all  the  expansion  is 
made  on  the  casing  side  of  the  blades  the  expansion  of  the  steam 
will  increase  the  leakage  around  the  tips  of  the  blades  next  to  the 
rotor  without  materially  affecting  the  leakage  at  the  tips  nearest 
the  casing. 

The  leakage  of  steam  around  the  tips  of  the  blades  depends, 
of  course,  again  upon  the  amount  of  the  radial  clearance.  Im- 
provement in  the  design  of  reaction  turbines  will  be  largely 
accomplished  (i)  by  skillful  designing  and  machine  work  to 
permit  the  reduction  of  radial  clearances  and  (2)  by  increasing 
the  blade  speed.  In  fact  the  question  of  allowable  radial  clear- 
ance depends  finally  on  the  blade  speed.  If  the  blade  speed  is 
increased  it  is  possible  to  use  higher  steam  velocities  with  larger 
pressure  drop  per  stage,  and  consequently  fewer  stages.  This 
is  apparent  also  from  an  inspection  of  the  designing  formula  on 
page  97.  With  fewer  stages  a  shorter  rotor  is  required  which 
will  also  be  proportionately  stiff  er;  and  with  a  stiff  shaft  it  is  pos- 
sible to  allow  very  small  radial  clearances,  provided,  of  course, 
temperature  effects  are  carefully  studied.  On  the  other  hand, 
by  increasing  the  pressure  drop  per  stage  the  tendency  for  leakage 


I08  THE    STEAM   TURBINE 

is  increased,  but  there  is  also  a  compensating  effect  in  that  the 
number  of  leakage  areas  is  correspondingly  reduced. 

The  reader  will  have  observed  that  the  design  of  reaction 
turbines  is  largely  by  "  cut  and  try  "  methods.  For  this  reason 
it  is  a  financial  absurdity  for  a  manufacturer  to-day  to  begin 
making  reaction  turbines  unless  he  has  practically  unlimited 
resources  and  can  obtain  from  makers  of  similar  machines  at 
not  too  large  a  cost  the  results  of  their  experiences. 

The  method  explained  here  of  determining  the  important  and 
unique  parts  in  the  design  of  a  reaction  turbine  for  a  given  set 
of  conditions,  as  regards  maximum  output,  steam  consumption, 
pressure,  superheat  and  vacuum,  although  very  simple  in  all 
essentials  as  regards  standard  practice,  gives  results  on  which  it 
is  impossible  to  improve  by  the  most  elaborate  mathematical 
analysis  imaginable.  In  fact  all  elaborately  mathematical  anal- 
yses of  the  action  of  steam  in  a  reaction  turbine  depend  finally 
on  the  substitution  of  certain  coefficients,  most  of  which  have 
no  basis  in  fact. 


DESIGN  OF  A   COMBINED   IMPULSE  AND   REACTION 
TURBINE. 

The  design  as  regards  the  general  method  for  a  combined 
impulse  and  reaction  turbine  will  be  similar  to  that  for  the 
design  of  the  impulse  turbine  (pages  86-95),  which  had  two 
velocity  stages  in  the  first  pressure  stage  with  two  rows  of  moving 
blades  or  buckets.  All  the  other  pressure  stages  had  only  one 
velocity  stage  and  therefore  only  one  row  of  moving  blades. 
For  the  combined  impulse  and  reaction  (similar  to  Fig.  107, 
page  174),  the  first  stage  might  well  be  arranged  with  two  velocity 
stages  as  in  the  design  referred  to,  and  the  other  stages  could  then 
be  designed  as  a  separate  two  section  reaction  turbine,  assuming 
the  steam  to  enter  the  reaction  portion  of  the  turbine  at  the 
quality  determined  by  adiabatic  expansion  in  the  first  stage. 
Assume  twice  as  much  work  is  done  in  the  second  reaction  sec- 
tion as  in  the  first.  With  this  understanding  it  is  certainly 


STEAM   TURBIXE   TYPES   AXD    BLADE    DESIGN  Io8a 

unnecessary  to  go  through  again  in  detail  the  details  of  the  de- 
signing of  the  blading  of  the  reaction  portion. 

Another  method  would  be  to  divide  up  the  work  to  be  done 
by  the  various  sections  as  in  the  design  of  the  complete  reaction 
turbine;  that  is,  one-fourth  of  the  work  would  be  done  by  the 
first  pressure  stage  having  its  two  velocity  stages  (as  in  Fig.  107, 
page  174),  another  fourth  would  be  done  by  the  first  section  of 
the  reaction  blading,  and  the  remaining  half  of  the  work  by  the 
second  and  last  section  of  reaction  blading. 

In  the  case  of  a  double-flow  turbine  in  which  the  low-pressure 
section  is  divided  into  two  halves,  the  equation  given  on  page 
lood  for  minimum  permissible  diameter  of  the  last  stage  would 
be  found  to  be  approximately 

d22  =  |  X  .57  X  output  in  kilowatts 
or  dz2  =  .285  X  output  in  kilowatts. 

In  the  same  connection  purchasers  of  steam  turbines  should 
guard  well  their  interests  by  exercising  good  business  judgment 
in  purchases.  Like  all  other  kinds  of  machinery,  there  will  be 
"  troubles  "  with  new  types  of  steam  turbines,  and  unless  the 
manufacturer  is  known  to  be  financially  responsible  and  well 
established  in  the  business,  the  purchaser  should  not  buy  until 
he  has  made  very  careful  investigations  of  the  merits  of  the  new 
machines;  and  he  should  always  insist  on  having  accurate  and 
complete  acceptance  tests,  made  preferably  by  disinterested 
engineers  of  repute. 

Exercise.  —  Design  the  blades  for  a  3Oo-horsepower  (maxi- 
mum output)  impulse  turbine  with  two  pressure  stages  and  two 
velocity  stages  in  each  pressure  stage  (Curtis  type).  Initial  ad- 
mission pressure  is  165  pounds  per  square  inch  absolute  at  100°  F. 
superheat,  and  the  exhaust  pressure  is  i  pound  per -square  inch 
absolute.  Blade  speed  500  feet  per  second.  Reheating  factor 
is  50  per  cent.  Use  8  nozzles  and  arrange  for  equal  energy  dis- 
tribution in  the  various  stages.  Nozzle  loss  is  2  per  cent,  of 
velocity,  and  take  blade  losses  from  curves  on  page  85. 


Io8b  THE   STEAM   TURBINE 

Exercise.  —  Design  of  the  blades  for  a  reaction  turbine  with 
50  stages  (Parsons  type)  for  the  same  conditions  of  power,  pres- 
sures and  superheat  as  in  the  preceding  example. 

Exercise.  —  Design  the  blades  of  a  combined  impulse  and  re- 
action turbine,  having  a  single  pressure  stage  of  the  impulse 
type  with  two  velocity  stages  (Curtis  type)  and  the  usual  type 
of  reaction  blading.  Conditions  of  power,  pressures  and  super- 
heat are  to  be  the  same  as  in  the  preceding  exercises. 


GENERAL  COMPARISON  OF  COMMERCIAL  IMPULSE  AND 
REACTION  TURBINES. 

IMPULSE. 

1.  Few  stages. 

2.  Expansion  in  nozzles. 

3.  Large  drop  in  pressure  in  a  stage. 

4.  Initial  steam  velocities  are  in  general  high  (1000  to  4000  feet  per  second). 

5.  Blade  velocities  400  to  1200  feet  per  second. 

6.  Best  efficiency  when  blade  velocity  is  nearly  half  the  initial  velocity  of  the 
steam.     For  a  single  wheel  per  pressure  stage. 

REACTION. 

1.  Many  stages. 

2.  No  nozzles. 

3.  Small  drop  in  pressure  in  a  stage. 

4.  All  steam  velocities  are  low  (30x5  to  600  feet  per  second). 

5.  Blade  velocities  150  to  400  feet  per  second. 

6.  Best  efficiency  when  the  blade  velocity  is  nearly  equal  to  the  highest  velocity 
of  the  steam. 


STEAM  TURBINE  TYPES  AND   BLADE  DESIGN         109 


The  work  done  per  stage  is  always  much  greater  in  current 
practice  in  impulse  than  in  reaction  turbines.  For  the  same 
total  limits  of  pressure  the  work  per  stage  is  inversely  propor- 
tional to  the  number  of  stages. 

In  general,  we  may  say  that  mechanical  considerations  and 
the  speed  at  which  machinery  can  be  conveniently  operated 
determine  the  size  and  number  of  revolutions  at  which  a  tur- 
bine can  be  run.  In  a  good  design  about  the  same  total  efficiency 
is  obtained,  whether  the  turbine  is  classified  as  an  impulse  or  a 
reaction  machine. 

Radial  Blade  Clearances.  In  impulse  turbines  the  radial 
clearance  (between  the  blade  ring  and  the  inside  of  the  casing) 


FIG.  57.    Illustrates  Radial  and  Axial  Clearances  in  an  Impulse  Turbine. 

is  not  important.  It  is  one  of  the  first  principles  of  a  good 
design  of  an  impulse  turbine  that  the  blades  shall  be  made  long 
enough  to  allow  the  steam  to  be  discharged  through  them  freely 
without  "choking"  the  flow  and  "spilling"  steam  over  the  outer 
edges  of  the  blades.  Since  the  pressure  is  the  same  on  the  two 
sides  of  the  blades,  radial  blade  clearances  in  impulse  turbines 
can  be  made  of  generous  dimensions.  (See  Figs.  57  and  119, 
in  which  Curtis  designs  are  shown.) 

In  reaction  turbines,  on  the  other  hand,  it  is  very  necessary  to 


110  THE   STEAM   TURBINE 

make  radial  clearances  as  small  as  is  mechanically  possible, 
because  in  these  turbines  the  steam  expands  in  the  moving  as 
well  as  in  the  stationary  blades  and  there  is  a  drop  in  pressure 
between  the  two  sides  of  every  row  of  blades.  On  account  of 
this  pressure  drop  there  is  a  continuous  flow  of  steam  around  the 
edges  of  the  blades,  which  is  large  or  small  in  amount  in  pro- 
portion to  the  size  of  the  radial  clearances.  The  clearance 
between  the  stationary  blades  fixed  to  the  casing  and  the  surface 
of  the  rotor  is  of  course  just  as  important  as  that  between  the 
moving  blades  and  the  casing.  An  American  manufacturer  of 
the  Parsons  reaction  turbines  states  that  the  radial  clearances  are 
from  .02  to  .10  inch,  varying  with  the  diameter  of  the  drum. 
These  limits  are  given  for  drums  between  i  foot  and  10  feet  in 
diameter.  Radial  clearances  of  large  sizes  of  Parsons  turbines 
made  by  Brown-Boveri  &  Co.  are  from  2  to  3  millimeters 
(.08  to  .12  inch).  Attainment  of  minimum  safe  radial  clear- 
ances is  the  goal  for  every  designer  of  reaction  turbines. 

Axial  Blade  Clearances.  Axial  clearances  with  respect  to 
impulse  and  reaction  turbines  present  conditions  just  opposite 
from  those  for  radial  clearances.  In  reaction  turbines,  axial 
clearance  is  not  an  important  factor  in  the  design.  Until  recently, 
however,  it  was  considered  very  important  in  the  design  of  im- 
pulse turbines  to  make  the  axial  clearance  between  the  moving 
blades,  and  nozzles  or  stationary  blades,  as  small  as  possible;  and 
indeed,  unfortunately,  some  impulse  turbines  in  the  early  days 
were  built  with  very  small  axial  clearances,  so  that  the  least 
vibration  of  the  shaft  would  cause  striking  of  the  moving  blades 
against  the  nozzles.  It  has  been  shown,  however,  by  actual 
experience  as  well  as  by  experiment  that  axial  clearances  can  be 
made  as  large  as  .20  inch  without  appreciable  loss;  or,  in  other 
words,  practically  as  large  as  in  reaction  turbines  —  usually 
about  .10  to  .20  inch. 

The  difficulties  of  the  designers  of  the  first  commercial 
impulse  turbines  can  well  be  imagined  when  it  was  considered 
so  essential  to  make  the  axial  clearances  not  more  than  .02  or  .03 
inch.  In  the  case  of  one  small  turbine  built  with  three  stages 


STEAM  TURBINE  TYPES  AND  BLADE   DESIGN         III 

the  axial  expansion  of  the  shaft  in  the  length  included  between 
the  high  pressure  nozzle  mouths  and  the  blades  of  the  third 
stage  was  .10  inch  by  actual  measurement.  To  allow  for  a  shift- 
ing of  the  blades  of  .10  inch  with  only  .03  inch  axial  clearance  in 
a  turbine  with  velocity  stages  was  not  an  easy  problem. 

Axial  clearances  in  Curtis  impulse  turbines  are  .06  to  .15  inch 
for  soo-kilowatt  sizes,  and  in  larger  machines  are  sometimes  as 
much  as  .25  inch.  In  Rateau  impulse  turbines  these  clearances 
are  from  .12  to  .25  inch.* 

Materials  for  Blades  and  Erosion.  Rolled  steel  is  a  very 
suitable  metal  for  turbine  blades  when  used  for  dry  or  super- 
heated steam  at  either  high  or  low  velocities  if  the  turbine  is 
kept  in  practically  continuous  operation.  Wet  steam,  however, 
will  wear  away  steel  blades  very  rapidly  by  erosion,  and  when  a 
turbine  fitted  with  steel  blades  is  idle  for  days  at  a  time  the 
blades  will  corrode,  so  that  when  it  is  started  again  the  particles 
of  iron  oxide  (rust),  will  be  carried  away  by  the  steam  to  act 
like  a  sand  blast  on  the  blades  in  succeeding  stages.  Steel  is 
an  exceptionally  good  material  for  blades  under  favorable  con- 
ditions because  it  can  be  rolled  cheaply  into  bars  of  any  shape 
of  section,!  and  it  is  unequaled  for  strength.  Copper  alloys, 
known  in  the  trades  as  " extruded  metal,"  are  made  into  bars 
of  any  shape  of  section  by  " drawing"  as  wire  is  manufactured. 

No  metal  has  all  the  physical  properties  desirable  in  a  blading 
material.  Recently  a  compound  metal  known  as  Monnot  or 
"duplex"  metal  has  been  developed.  It  consists  of  a  steel  core 
covered  with  a  thin  copper  sheathing  chemically  .welded  to  the 
steel  in  such  a  perfect  manner  that  the  blades  may  be  drawn 
cold  from  the  original  ingot  into  the  required  finished  section 

*  In  impulse  turbines  with  nozzles  discharging  radially  into  blades  or  buckets 
on  the  rim  like  the  Sturtevant,  Terry,  or  Riedler-Stumpf  types,  it  is  stated  that  there 
is  no  appreciable  change  in  velocity  loss  when  the  radial  clearance  (between  the 
nozzle  and  the  buckets)  is  increased  from  .10  to  .40  inch. 

f  Rolled  bars  are  cut  up  into  lengths  corresponding  to  the  height  of  the  blade 
plus  an  additional  length  for  dovetailing  into  the  rim  of  the  turbine  wheel.  When 
this  dovetailing  method  is  used  (Fig.  63)  the  blades  are  separated  from  each  other 
by  "  spacing  pieces  "  of  suitable  shape  to  fit  between  the  blades. 


112 


THE  STEAM  TURBINE 


without  in  any  way  affecting  the  bond  between  the  copper  and 
the  steel.  Fig.  58  shows  an  etched  section  of  a  blade  of  this 
material  from  a  Westinghouse  turbine. 

Blades  like  those,  for  example,  in  Figs.  59-61,  which  are  too 
irregular  to  be  rolled  or  drawn  are  usually  cast  of  bronze  or  copper 


FIG.  58.     Etched  Section  of  a  Blade  made  of  Monnot  Metal  (Steel  and  Copper). 

alloys.  Forked  blades  (Fig.  59)  similar  to  those  used  in  Wilkin- 
son turbines  are  commonly  cast  with  the  forks  far  enough  apart 
so  that  they  will  pass  over  the  enlarged  section  of  the  rim  and  are 
forced  together  when  they  are  in  place.  Another  method  is  to 


FIG.  59.  FIG.  60.  FIG.  61. 

Designs  of  Steam  Turbine  Blades. 

cut  away  the  enlarged  part  of  the  rim  section  for  a  short  length, 
and  blades  cast  with  the  forks  in  their  normal  position  can  be 
inserted  at  this  place  and  can  then  be  pushed  around  on  the  rim 
till  all  the  blades  are  in  place.  The  parts  of  the  rim  cut  away 
must  be  replaced  to  secure  the  blades  at  that  section. 

The  blades  of  small  sizes  of  Curtis  turbines  are  sometimes  cut 
in  the  rim  of  a  solid  disk  by  automatic  machinery.     De  Laval 


STEAM  TURBINE  TYPES  AND  BLADE  DESIGN         113 


FIG.  63.     Dovetailed  Type  of  Blade  (Curtis). 


FIG.  64.     Typical  De  Laval  Blades. 


THE  STEAM  TURBINE 


blades  are  made  of  steel  forged  into  the  peculiar  shape  required 
for  insertion  into  the  disk  wheel.  (See  Fig.  64. ) 

It  is  stated  that  the  usual  alloy  used  in  England  for  blades  of 
Parsons  turbines  is  63  Cu  +  37  Zn;  but  any  zinc  alloy  is  quite 
unsuitable  for  superheated  steam  or  for  high  velocities. 

Fig.  65  *  shows  the  effect  of  the  erosion  due  to  steam  on  blades 


FIG.  65.     Photograph  of  Turbine  Blades  Showing  Erosion. 

made  of  Delta  metal  about  60  Cu  +  37  Zn  +  3  Fe.  These 
blades  were  held  stationary  in  a  steam  jet  for  128  hours.  The 
blades  on  the  left  side  of  the  figure  were  subjected  to  steam  at 
2900  feet  per  second;  and  those  on  the  right  to  steam  at  600 
feet  per  second.  Low-velocity  steam  eroded  the  blades  so  little 
that  the  tool  marks  put  in  the  blades  when  they  were  made  are 
still  visible. 

*  The  author  is  indebted  to  Mr.  Francis  Hodgkinson  for  this  photograph. 


CHAPTER    V. 
MECHANICAL  LOSSES  IN  TURBINES. 

IN  the  designs  of  turbines  on  the  preceding  pages  the  nozzle 
and  blade  efficiency  was  first  calculated,  and  then  the  total,  or 
"over-all,"  shaft  efficiency  was  obtained  by  subtracting  other 
losses  as  follows: 

(1)  Disk  and  blade  friction,  or  windage,  due  to  rotation  in  a 
fluid  medium  (steam). 

(2)  Leakage    of    the    steam    chiefly    through    the    clearance 
between  the  shaft  and  the  diaphragms  of  a  multi-stage  impulse 
turbine  and  through  the  radial  blade  clearances  in  a  reaction 
turbine. 

(3 )  Bearing  and  stuffing-box  friction  losses. 

(4)  Radiation. 

Of  these  the  first  three  are,  in  a  way,  mechanical  losses  in  the 
sense  that  the  details  of  mechanical  design  largely  determine 
their  values. 

The  first  of  these  losses,  disk  and  blade  rotation  loss,  is  by  far 
the  most  important  and  will  be  discussed  first. 

Losses  Due  to  Friction  of  Turbine  Wheel  Revolving  in  Steam. 
Losses  due  to  revolving  disks  or  wheels  in  steam  are  very  diffi- 
cult to  determine  with  accuracy.  Tests  to  determine  these 
losses  are  usually  made  with  the  wheel  rotating  in  stagnant 
steam,  and  it  is  practically  impossible  to  have,  under  these  con- 
ditions, steam  of  the  same  quality  or  superheat  in  all  parts  of  the 
casing.  A  number  of  formulas  have  been  proposed  for  the 
friction  losses  of  disks  and  blades  in  dry  saturated  steam,  but 
there  is  no  good  agreement  of  the  results  of  different  experi- 
menters. In  fact  no  great  accuracy  can  be  expected  because 
there  is  no  doubt  that  the  exponents  of  logarithmic  friction 

"5 


IX6  THE   STEAM  TURBINE 

curves  plotted  from  such  tests  vary  considerably  with  the  details 
of  design,  and  besides,  it  is  very  difficult  to  get  good  tests.* 

An  important  reason  why  the  tests  from  different  designs  of 
turbines  do  not  agree  better  is  that  clearances  between  moving 
and  stationary  parts  have  an  appreciable  effect.  If  the  clear- 
ances all  around  the  wheel  are  very  small  the  wheel  and  blade 
friction  loss  will  be  somewhat  less  than  for  a  wheel  revolving  in 
large  clearance  spaces.  This  effect  is  most  marked  at  low  speeds. 
When  higher  speeds  are  reached  there  is  more  tendency  for  the 
wheel  to  "cut  through"  the  surrounding  steam  without  increas- 
ing the  "  disturbance"  in  proportion  to  the  increase  in  speed. 

The  author  has  from  time  to  time  investigated  large  numbers 
of  tests  to  determine  the  friction  losses  of  wheels  and  blades  of 
turbines  in  steam  and  air,  and  this  experience  has  shown  that  the 
following  formulas  will  give  fair  average  results  for  forward 
running  in  practically  stagnant  steam.  The  rotation  loss  or 
skin  friction  of  a  plain  disk  |  revolving  in  dry  saturated  steam  is 
expressed  by  the  following  formula  in  horsepower  : 

(25) 

where  d  is  the  diameter  of  disk  to  inner  edge  of  blade  in  feet. 
u  is  the  peripheral  velocity  of  disk  in  feet  per  second.  J 
y  is  the  density  of  surrounding  medium  in  pounds  per 

cubic  foot  (reciprocal  of  the  specific  volume). 
A  similar  term  to  determine  the  rotation  loss  of  one  row  of 
blades  F6  (without  the  disk),  in  horsepower,  is 

,  (26) 


ioo 


*  The  peculiar  circumstance  that  water  in  the  liquid  state  can  exist,  indefinitely, 
in  the  presence  of  superheated  steam,  leading  some  to  propose  a  vergasungs- 
•wdrme,  is  one  of  the  greatest  difficulties. 

f  Similar  to  those  in  Curtis  and  Rateau  turbines.  On  account  of  the  thick 
hubs  of  De  Laval  disks  (Figs.  83  and  84),  about  15  per  cent,  should  be  added  to 
the  results  given  by  equation  (25)  to  allow  for  the  larger  surface  of  these  disks. 

J  It  is  often  stated  that  the  disk  and  blade  friction  losses  vary  as  the  third  power 
of  the  speed.  But  this  value  cannot  be  stated  with  any  claim  to  great  accuracy. 
Experimenters  do  not  all  agree  on  this  value,  and  values  from  2.5  to  3.5  are  given  by 
different  authorities.  The  author,  from  the  result  of  the  experiments  he  has 


MECHANICAL  LOSSES  IN  TURBINES  117 

where  1  =  length  of  blades  in  inches  excluding   the   band    (if 
there  is  one),  and  d,  u,  and  y  are  used  as  before. 

For  a  simple  turbine  wheel  with  only  one  row  of  blades  we 
can  write  for  the  total  rotation  loss  F,  in  horsepower : 

F,  =  (.o8d  +  .Sl1'5)^  — T'8dy.  (27) 

(  IOO  ) 

The  density  of  superheated  steam  varies  with  the  amount  of 
superheat,  so  that  by  adding  the  following  notation, 

yd  =  density  of  dry  saturated  steam  at  the  pressure  of  the 

surrounding  medium  in  pounds  per  cubic  foot, 
D  =  superheat  in  degrees  F., 
vd  =  specific  volume  of  dry  saturated  steam  at  the  pressure  in 

the  surrounding  medium  in  pounds  per  cubic  foot, 
and  using  the  following  equation  for  specific  volume  v«  of  super- 
heated steam  given  on  page  53, 

•  v5  =  (i  +  .00065  D)2  vrf, 

we  have  the  following  formulas,  taking  the  place  of  (25),  (26), 
and  (27)  above,  for  superheated  steam: 

/  u  V8 

( )     7d 

Tw  =  .08  d2          ™Q6    p)2  '  (25' ) 

i^\ 


F,  -  (.08  d  +  .3  I1'5)        XI°°'     74—  •  (27') 

(i  +  .00065  D)2 

investigated,  considers  the  2.8  power  a  good  average  value  suitable  for  practically  all 
conditions.  In  the  value  of  the  exponent  this  rotation  loss  resembles  train  and 
ship  resistance.  The  windage  loss  of  dynamos  properly  designed  for  high  speeds  is 
a  curve  of  the  second  power.  When  the  windage  loss  curve  of  a  dynamo  shows  an 
exponent  of  3  or  3.5  it  must  be  inferred  that  the  machine  was  not  properly  designed 
for  high  speeds.  It  may  be  interesting  to  the  practical  men  reading  this  book  to  know 
how  the  exponent  is  obtained  from  a  test.  This  is  done  most  conveniently  by 
plotting  on  any  suitable  coordinate  paper  the  logarithms  of  the  loss  for  the  ordinates, 
and  the  logarithms  of  the  speed  for  the  abscissas.  The  tangent  of  the  curve  is  the 
value  of  the  exponent  if  the  scales  of  ordinates  and  of  abscissas  are  the  same. 


n8 


THE   STEAM  TURBINE 


Or  the  curve  given  in  Fig.  68  can  be  used  to  correct  equations 
(25),  (26),  and  (27)  by  means  of  a  coefficient. 

While  the  effect  of  superheating  is  to  reduce  these  losses, 
moisture,  on  the  other  hand,  increases  them  very  appreciably. 


.70 


40        60 


100      120      140      160      180      200      220      240 
Superheat-Degrees  Fahr. 


FIG.  68.    Curve  to  Correct  Rotation  Losses  for  Superheat. 


2.60 
2.40 
2.20 
2.00 
1.80 
1.60 
1.40 
1.20 
1.00 

/ 

/ 

/ 

i 

/ 

/ 

/ 

/ 

/ 

/ 

/ 

/ 

/ 

/ 

X 

x 

A 

!/ 

^ 

s* 

^ 

-"'' 

e== 

'—  " 

p-^-  - 

****** 

2          4 


Percent.  Moisture  in  Steam 


FIG.  69.     Curve  to  Correct  Rotation  Losses  for  Moisture. 

Fig.  69  shows  a  curve  giving  the  coefficients  to  be  applied  to  the 
losses  calculated  by  the  above  formulas  for  dry  saturated  steam 
to  correct  for  moisture. 

Example.     Calculate  the  frictional  rotation  loss  of  a  disk  3 
feet  in  diameter  of  a  non-condensing  single  stage  turbine  (steam 


MECHANICAL  LOSSES  IN  TURBINES  119 

pressure  15  pounds  per  square  inch  absolute)  when  the  steam 
is  (i)  dry  saturated,  (2)  superheated  100°  F.,  (3)  10  percent 
wet.  The  speed  is  3600  revolutions  per  minute.  Determine 
also  the  rotation  loss  of  a  single  row  of  blades  i  inch  long  on 
this  disk. 

At  high  peripheral  speeds  the  rotation  loss  of  a  non-condensing 
turbine  with  the  wheels  revolving  in  steam  at  atmospheric  pres- 
sure is  quite  large,  as  the  example  above  illustrates.  This  loss 
decreases,  however,  very  rapidly  with  increasing  vacuum,  and 
is,  in  fact,  nearly  proportional  to  the  pressure.  This  fact  is 
not,  however,  always  appreciated  by  designers.  Of  course, 
when  disk  and  blade  rotation  losses  are  being  calculated  for  a 
series  of  pressures  for  the  several  stages  of  a  turbine,  as  is  usually 
done  before  deciding  on  the  nozzle  proportions,  it  is  only  neces- 
sary, if  the  wheel  dimensions  are  constant,  to  calculate  for  one 
pressure  and  determine  the  values  for  the  other  stages  by  multi- 
plying by  a  constant  representing  the  ratio  of  the  densities.  Of 
all  the  variables  in  equations  (25),  (26),  and  (27),  the  density  is 
the  only  term  varying  as  the  first  power.  For  most  work  it  will 
be  allowable  to  assume,  within  a  small  range,  the  density  pro- 
portional to  the  pressure ;  that  is,  if  the  disk  and  blade  loss  has 
been  calculated  in  steam  at  some  given  pressure,  the  correspond- 
ing friction  loss  at  any  other  pressure  may  be  found  by  the  ratio 
of  the  pressures. 

The  disk  and  blade  rotation  losses  of  a  Parsons  or  other  drum 
type  may  be  calculated  with  the  above  formulas  by  calcu- 
lating the  loss  for  each  group  of  blades  of  the  same  length  and 
diameter  and  adding  to  the  sum  of  the  blade  losses  the  rotation 
loss  due  to  disks  approximately  equivalent  to  the  outside  surface 
of  the  drum.  As  the  friction  loss  due  to  the  drum  itself  is  small 
compared  with  that  of  the  many  rows  of  blades,  no  great  accuracy 
need  be  attempted  in  this  calculation. 

In  small  sizes  of  steam  turbine-generators  the  rotation  loss  is  a 
considerable  percentage  of  the  total  output.  The  disk  and  blade 
loss  of  a  single  stage  turbine  with  a  single  row  of  blades,  rated  by 
the  manufacturer  at  about  250  kilowatts  at  3600  r.p.m.,  is  shown 


120 


THE  STEAM  TURBINE 


in  Fig.  70.  The  curves  show  that  the  rotation  or  windage  loss 
of  the  generator  alone  is  about  30  kilowatts  and  the  total  rotation 
loss  is  50  kilowatts  or  20  per  cent,  of  the  rated  output.  Similarly 


2000 


FIG.  70.     Rotation  Loss  Curves  of  25o-Kilo\vatt  Turbine-Generator. 

the  total  rotation  loss  of  a  2000  to  3000  kilowatt  turbine-gen- 
erator is  from  10  to  15  per  cent,  of  the  rated  output. 

Method  of  Making  Tests  to  Determine  Wheel  and  Blade 
Rotation  Losses  of  a  Steam  Turbine.  The  simplest  method  for 
making  such  a  test,  and  the  one  commonly  employed,  is  to  attach 
an  electric  motor  to  the  turbine  shaft  (sometimes  in  a  direct- 
connected  set  the  generator  is  used  as  a  motor)  and  run  it  at  a 
number  of  different  speeds.  In  taking  a  series  of  speeds,  no 
observations  are  made  until  conditions  have  become  "  steady, " 
and  the  speed  must  be  held  constant  for  several  minutes  so  that 
a  number  of  readings  can  be  taken  on  the  electrical  instruments 
measuring  the  input  of  the  motor.  The  results  give  the  rota- 
tion loss  of  the  wheel  and  blades  in  steam  as  well  as  bearing 
friction  and  the  rotation  or  "windage"  and  electrical  losses  of  the 
motor.  Then  the  turbine  wheel  is  removed,  leaving  the  packing 


MECHANICAL  LOSSES   IN   TURBINES  121 

at  the  generator  end  of  the  turbine  on  the  shaft,  and  the  motor  is 
run  alone.  The  power  now  measured  is  that  required  to  over- 
come the  rotation  and  electrical  losses  of  the  generator  and  the 
bearing  friction.  Curves  of  power  and  speed  as  variables  (Fig. 
70)  are  plotted  for  each  set  of  observations,  and  the  disk  arid 
blade  loss  is  determined  by  subtracting  the  ordinates  of  one  curve 
from  those  of  the  other.  It  may  be  assumed  with  sufficient 
certainty  that  the  weight  of  the  turbine  wheel  itself  would  not 
alter  the  bearing  losses  to  any  considerable  extent.* 

The  important  fact  that  all  results  given  here  are  for  disks  and 
blades  revolving  in  a  stagnant  medium  must  not  be  overlooked, 
and  it  must  not  be  assumed  that  the  results  will  be  the  same 
under  actual  operating  conditions.  It  may  be  a  coincidence  that 
the  losses  are  the  same  in  both  cases.  Under  operating  condi- 
tions, the  spaces  between  the  wheel  blades  are  filled  with  steam 
flowing  from  the  nozzle  over  the  blades  and  then  to  the  condenser. 
Now  it  has  been  shown  by  a  series  of  experiments  by  Laschef 
of  the  Allgemeine  Electricitat  Gesellschaft  (Berlin)  that  increasing 
the  number  of  nozzles  around  the  turbine  wheel  reduces  the  disk 
and  blade  rotation  losses.  These  losses  in  the  blades  are  very 
largely  due  to  the  fan  action  of  the  blades  which  start  currents  of 
steam  just  as  a  centrifugal  fan  does.  In  other  words,  this  is  what 
Stodola  calls  "ventilation."  With  steam  flowing  through  the 
blades,  this  fan  action  is  largely  prevented  and  the  losses  are  con- 
sequently reduced.  Another  reason  why  the  disk  and  blade 
rotation  losses  should  be  less  when  the  turbine  is  operating  than 
they  are  in  stagnant  steam,  is  that  they  are  really  friction  losses, 
or  a  conversion  of  kinetic  energy  into  heat,  with  the  effect  of 
either  superheating  or  drying  the  steam.  In  a  turbine  with 
more  than  one  stage  a  part  of  the  heat  energy  gained  as  the  result 
of  the  friction  is  converted  in  the  next  expansion  into  kinetic 
energy  or  velocity.  It  is  usually  assumed  that  about  15  percent. 

*  It  may  be  interesting  to  observe  that  since  disk  and  blade  friction  is  pro- 
portional to  the  density  of  the  medium,  the  friction  is  therefore  greater  in  air  than 
in  dry  saturated  steam  at  atmospheric  pressure.  This  is  shown  by  experiments 
published  by  Lewicki  in  Zeit.  Verein  deutscher  Ingenieure,  March  28,  1903. 

|  Stodola,  Die  Dampfturbinen,  third  edition,  page  130. 


122  THE    STEAM  TURBINE 

of  the  disk  and  blade  losses  are  regained  by  the  reheating,  and 
that  therefore  the  actual  friction  losses  in  an  operating  turbine 
are  about  this  amount  smaller  than  in  stagnant  steam.  In  cases 
of  full  admission  true  blade  friction  disappears ;  and  a  proportion- 
ate reduction  will  also  take  place,  according  to  the  degree  of 
admission,  when  it  is  partial. 

Investigation  of  wheel  and  blade  friction  losses  by  the  author, 
using  a  modification  of  the  method  first  suggested  by  Lasche  of 
Berlin,  did  not  show  the  reduction  in  these  losses  to  be  expected 
when  determined  under  operating  conditions.  These  results, 
however,  cannot  be  considered  conclusive,  as  the  type  of  machine 
used  was  not  well  suited  for  the  purpose,  and  only  25  per  cent,  of 
the  blades  were  filled  with  steam.  It  has  been  stated  that  when 
a  large  quantity  of  steam  passes  into  the  casing  through  a  suit- 
able opening  without  passing  through  nozzles  and  escapes  through 
the  exhaust  (without  increasing  the  pressure),  the  disk  and  blade 
rotation  losses  are  increased  as  much  as  20  per  cent.  This 
apparently  is  an  influence  to  counteract  the  effect  of  filling  the 
blades. 

In  all  the  analysis  that  has  preceded  there  are  so  many  uncer- 
tain variables  entering  that  it  is  impossible  to  get  agreement, 
although,  apparently,  we  have  a  large  amount  of  data  from  which 
to  draw.  It  may  be  stated,  however,  that  all  in  all,  the  best  data 
on  disk  and  blade  friction  seem  to  show  that  it  is  smaller  and  of 
less  significance  than  the  results  of  most  investigators  would 
show. 

A  little  space  should  be  given  to  Lasche's  very  interesting 
method.*  A  turbine-generator  set  was  used  in  which  the  number 
of  nozzles  discharging  into  the  turbine  could  be  regulated  and  the 
output  of  the  generator  was  observed  for  each  setting  of  valves, 
and  tests  with  varying  loads  were  made  at  a  number  of  different 
speeds.  The  turbine  wheel  was  then  removed  from  the  shaft, 
and  by  running  the  generator  as  a  motor  the  friction  losses  in  the 
stuffing-box  at  the  generator  end  of  the  turbine  and  in  the  bearings, 
as  well  as  the  windage  loss  of  the  generator,  were  determined. 

*  Stodola,  Die  Dampfturbinen,  third  edition,  page  131. 


MECHANICAL  LOSSES  IN  TURBINES 


123 


The  resistance  of  the  armature  and  brushes  was  also  measured  to 
calculate  the  heating  (Fr)  loss.  The  sum  of  these  losses  was  cal- 
culated for  a  number  of  loads  (kilowatts)  and  curves  similar  to 
those  in  Fig.  70  were  obtained.  Curve  A  in  Fig.  71  shows  the  elec- 


340 

I 

Ren 

320 

1 

| 

E 

ases 

/ 

280 

/ 

i 

260 
240 
220 
jjj   200 

I 

I 

1 

1 

/ 

1    % 

1 

/ 

I 

B 

A 

£» 

O   160 
140 
120 
100 
80 

1 

1 

/ 

1 

I 

1 

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f 

1 

A- 

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Output 

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n. 

40 

20 
0 

1 

/ 

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ii 

1 

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u 

7       1         2         3         4        5         6 

Nozzles  Open 

FIG.  71. 


FIG.  72. 


Curves  for  Determining  Disk  and  Blade  Rotation  Losses  at  Operating  Conditions. 

trical  output  at  3500  r.p.m.  Curve  B  in  the  same  figure,  represent- 
ing the  power  delivered  to  the  shaft  by  the  turbine,  was  obtained 
by  adding  to  the  generator  output  for  each  set  of  nozzles  open 
(curve  A)  the  corresponding  generator  losses  (windage,  heating, 
and  bearing  friction).  The  lower  portions  of  curves  A  and  B 


124  THE  STEAM  TURBINE 

are  practically  straight  lines,  and  by  producing  curve  B  to  the 
horizontal  axis,  its  intersection  represents  on  the  scale  of  abscissas 
the  disk  and  blade  rotation  losses  of  the  turbine  at  the  speed  of 
the  test  and  under  actual  operating  conditions. 

By  making  a  series  of  such  tests  at  different  speeds  curves  of. 
rotation  losses  can  be  made.  Fig.  72  shows  typical  curves  of  shaft 
output  for  speeds  of  3000,  3500,  and  4000  r.p.m.  Although 
this  method  requires  very  careful  experimenting,  the  same  must 
be  said  of  any  other  method  of  obtaining  these  losses,  and  most 
of  the  results  that  have  been  published  are  very  poor.  At  least 
it  must  be  admitted  that  by  this  method  a  number  of  uncertain 
factors  to  be  considered  in  the  "stagnant  steam"  method  are 
eliminated. 

The  lines  in  Fig.  72  are  really  the  same  as  "  Willans  lines"  and 
might  just  as  well  be  plotted  for  total  "flow"  of  steam  per  hour 
as  for  nozzles  open.  In  fact  in  turbines  where  there  are  no 
nozzles  the  "flow"  of  steam  must  be  used.  It  is  obvious  that 
any  load  curve  of  brake  horsepower  giving  the  total  steam  con- 
sumption can  be  used  to  determine  the  rotation  loss  by  producing 
the  "flow"  line  to  the  axis  on  which  the  output  is  scaled.  A 
good  check  on  the  results  of  such  rotation  loss  tests  is  secured  by 
observing  whether  the  lines  for  the  speeds  near  the  rated  speed 
cross  each  other  at  about  the  rated  output.  In  a  good  design  the 
speed-output  curve  will  be  like  the  curve  in  Fig.  80,  giving 
nearly  the  same  output  at  speeds  considerably  above  or  below 
the  rating. 

The  no  load  steam  consumptions  of  2000,  5000  and  9000  kilo- 
watt Curtis  turbine-generators  are  respectively  about  14,  12.5, 
and  8  per  cent,  of  that  at  full  load.  In  other  words  these 
percentages  are  only  from  one  to  two  per  cent,  greater  than  the 
sum  of  the  disk  and  blade  rotation  and  generator  windage 
losses.  Generator  windage  loss  is  probably  about  equal  to 
the  sum  of  all  the  turbine  losses.  It  is  generally  assumed  that 
the  no  load  steam  consumption  of  a  Parsons  turbine  (without 
the  generator)  is  about  12  per  cent,  of  that  at  the  normal 
maximum  output. 


MECHANICAL  LOSSES  IN  TURBINES  125 

It  is  stated*  that  at  no  load  the  steam  required  for  very  large 
reciprocating  engines  and  generators  is  probably  in  no  case  less 
than  15  per  cent,  of  that  used  at  full  load. 

Leakage  Loss.  The  other  important  mechanical  loss  in  a 
steam  turbine  is  that  due  to  the  leakage  of  steam  through  the 
passages  of  the  turbine  without  doing  work.  In  impulse  turbines 
of  more  than  one  stage  this  loss  is  chiefly  caused  by  the  leakage  of 
steam  between  the  shaft  and  the  diaphragms.  In  a  great  many 
turbines  no  satisfactory  packing  is  provided  at  these  places  and 
the  loss  is  sometimes  more  than  10  per  cent,  of  the  total  amount 
of  steam  supplied  to  the  turbine.  In  reaction  turbines  the  loss 
is  due  to  leakage  through  the  radial  clearance  passages  and  is 
large  or  small  in  proportion  to  the  size  of  these  clearances.  The 
loss  is  usually  assumed  to  be  about  5  per  cent,  in  good  Parsons 
turbines. 

Future  improvements  in  the  economy  of  all  types  of  steam  tur- 
bines will  depend  largely  on  the  success  of  designers  in  reducing 
these  leakage  losses.  For  impulse  turbines  an  improved  design 
has  been  patented  by  Wilkinson  (page  204).  In  reaction  tur- 
bines it  can  be  reduced  by  making  a  shorter  and  stiffer  shaft. 

Bearing  Friction.  This  loss  is  due  to  the  friction  of  the  shaft 
in  its  bearings,  and  in  a  De  Laval  turbine  the  friction  of  the  gears 
is  usually  included.  An  analysis  of  the  losses  in  a  De  Laval 
turbine  is  given  on  page  152,  where  the  bearing  friction  loss  is 
given  as  one  per  cent.  Bearing  friction  is  also  discussed  in  the 
footnote  on  page  92. 

*  Kruesi,  Proc.  Am.  Street  and  Interurban  Rail-way  Engineering  Association, 
1907. 


CHAPTER  VI. 
METHOD   FOR  CORRECTING  STEAM  TURBINE  TESTS. 

Standard  Conditions  for  Steam  Turbine  Tests.  If  tests  of  steam 
turbines  could  always  be  made  at  some  standard  vacuum,  super- 
heat, and  admission  pressure,  then  turbines  of  the  same  size  and 
of  the  same  type  could  be  readily  compared,  and  an  engineer 
could  determine  without  any  calculations  which  of  two  turbines 
was  more  economical  for  at  least  these  standard  conditions.  But 
steam  turbines  and  engines  even  of  the  same  make  are  not  often 
designed  for  and  operated  at  any  standard  conditions,  so  that  a 
direct  comparison  of  steam  consumptions  has  usually  no  signifi- 
cance. 

It  will  be  shown  now  how  good  comparisons  of  different  tests 
can  be  made  by  a  little  calculation  involving  the  reducing  of  the 
results  obtained  for  varying  conditions  to  assumed  standard 
conditions.  The  method  given  here  is  that  generally  used  by 
manufacturers  for  comparing  different  tests  on  the  same  turbine 
(a  "checking"  process)  or  on  different  types  to  determine  the 
relative  performance.  To  illustrate  the  method  by  an  applica- 
tion, a  comparatively  simple  test  will  first  be  discussed. 

*  Practical  Example.  Corrections  for  Full  Load  Tests.  The 
curve  in  Fig.  73  shows  the  steam  consumption  for  varying 
loads  obtained  from  tests  of  a  i25-kilowatt  steam  turbine 
operating  at  27.5  inches  vacuum,  50°  F.  superheat,  and  175 
pounds  per  square  inch  absolute  admission  pressure  (at  the 
nozzles).  It  is  desired  to  find  the  equivalent  steam  consump- 
tion at  28  inches  vacuum,  o°  F.  superheat,  and  165  pounds  per 
square  inch  absolute  admission  pressure  for  comparison  with 
•"guarantee  tests"  (Fig.  74)  of  a  steam  engine  of  about  the 
•same  capacity  operating  at  the  latter  conditions  of  vacuum, 
superheat,  and  pressure.  The  manufacturers  of  the  steam 

126 


METHOD  FOR  CORRECTING  STEAM  TURBINE  TESTS     127 


Steam  Consumption 
Lbs.Per  Kw.Hour 

S  g  8  8  8?  £  fe 

27.5  I  us.  Vacuum 
50°F  Superheat 
175  Lbs.Per  Sq.In.Abs. 
Pressure 

—  < 

\ 

\ 

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\ 

N 

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-*  — 

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-  — 

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•  —  — 

-    - 



- 

• 

•—     — 

r  ^ 

•^—  - 

)           20           40          60           80          100         120         140         .160         180        .200 
Output  in  Kilowatts 

FIG.  73.   Load  Curve  of  a  Typical  i25-Kilowatt  Steam  Turbine. 

35 


15 


28  Ins. Vacuum 

0°F  Superheat 

165  Lb&Per  Sq.In.Abs. 

Pressure 


A.  Steam  Consumption  of  Engine 

B.  Corrected  Steam  Consumption 
of  Steam  Turbine. 


20          40 


0  80          100         120         140         160         180         200 

Output  in  Kilowatts 

FIG.  74.    Comparative  Load  Curves  of  a  Reciprocating  Steam  Engine  and  a  Steam 
Turbine  —  Both  of  125  Kilowatts  Capacity  at  Full  Load 


128  THE  STEAM  TURBINE 

turbine  have  provided  the  curves  in  Figs.  75,  76,  and  77 
showing  the  change  of  economy  with  varying  vacuum,  superheat, 
and  pressure.  With  the  help  of  these  correction  curves,  the 
steam  consumption  of  the  turbine  can  be  reduced  to  the  conditions 
of  the  engine  tests.  Fig.  75  shows  that  between  27  and  28  inches 
yacuum  a  difference  of  one  inch  changes  the  steam  consumption 
i.o  pound.  Fig.  76  shows  a  change  of  2.0  pounds  per  100°  F. 
^superheat,  and  from  Fig.  77  we  observe  a  change  of  5.0  pounds 
:in  the  steam  consumption  for  100  pounds  difference  in  admission 
pressure.  Compared  with  the  engine  tests  the  steam  turbine  was 
operated  at  .5  inch  lower  vacuum,  50°  F.  higher  superheat,  and 
jo  pounds  higher  pressure.  At  the  conditions  of  the  engine 
tests,  then,  the  steam  consumption  of  the  steam  turbine  should  be 
.reduced  .5  pound  to  give  the  equivalent  at  28  inches  vacuum, 
but  is  increased  i.o  pound  to  correspond  to  o°  F.  superheat,  and 
,5  pound  more  to  bring  it  to  165  pounds  absolute  admission 
pressure.  The  full  load  steam  consumption  for  the  steam  tur- 
bine at  the  conditions  required  for  the  comparison  is,  therefore, 
24.5  —  .5  +  i.o  +  .5,  or  25.5  pounds.* 

Persons  who  are  not  very  familiar  with  the  method  of  making 
these  corrections  will  be  liable  to  make  mistakes  by  not  knowing 
whether  a  correction  is  to  be  added  or  subtracted.  A  little 
thinking  before  writing  down  the  result  should,  however,  prevent 
such  errors.  When  the  performance  at  a  given  vacuum  is  to  be 
corrected  to  a  condition  of  higher  vacuum  the  correction  must  be 
subtracted  because  obviously  the  steam  consumption  is  reduced 
by  operating  at  a  higher  vacuum.  When  the  steam  consumption 
with  superheated  steam  is  to  be  determined  in  its  equivalent  of 
dry  saturated  steam  (o°  superheat)  the  correction  must  be  added 
because  with  lower  superheat  there  is  less  heat  energy  in  the 
steam  and  consequently  there  is  a  larger  consumption.  Usual 

*  The  corrected  steam  consumption  is  found  to  be  nearly  the  same  as  that  which 
the  three  correction  curves  show  for  the  same  conditions,  that  is,  about  25.0  pounds. 
If  there  had  been  a  difference  of  more  than  about  5  per  cent,  between  the  corrected 
steam  consumption  and  that  of  the  correction  curves  for  the  same  conditions,  the 
"  ratio  "  method  as  explained  on  page  130  for  fractional  loads  should  have  been  used 
also  for  full  load. 


METHOD  FOR  CORRECTING  STEAM  TURBINE  TESTS     129 


0  F.Superheat 

165  Lbs.Per  Sq.Iu.Abs. 

Ful  ILoadC  125  Kw.) 


20  21.          22          23  .24          25          26          .27          28  29 

Vacuumrinches  of  Mercury  , 

FIG.  75.   Vacuum  Correction  Curve  for  a  i25~Ki!owatt  Steam  Turbine. 


Steam  Consumption 
Lbs.  per  Kw.Hour 

Z  S  8  %  S 

28  Ins.  Vacuum 
165Lbs.PerSq.In.Abs. 
FuHLoad(125Kw.) 

mmm 

Ci__ 

•  — 

—  —  . 

—  • 

•••  •• 

••  ••. 

—  —  . 

^  • 

—  — 

-  ..-• 

—  ^ 

—  • 

^       - 

—   •• 

—      •• 

—     

—      -. 

•        • 

—     •• 

>           90          40           60          80          100         120        140          1GO         130         200 
.  Superheat-Degs.Fahr. 
"IG.  76.    Superheat  Correction  Curve  of  a  125-Kilowatt  Steam  Turbine. 

28  Ins.  Vacuum 

0°F.  Superheat 

Full  Load ( 125  Kw.) 


100       110          120         130         140         150         160         1TO         180         190         200 

Steam  Pressure-Lbs.Per  Sq.In.Abs. 
FIG.  77.   Pressure  Correction  Curve  of  a  125-Kilowatt  Steam  Turbine. 


130  THE  STEAM  TURBINE 

corrections  for  differences  in  admission  pressure  are  not  large; 
but  it  is  well  established  that  the  economy  is  improved  by 
increasing  the  pressure. 

Corrections  for  Fractional  Loads.  It  is  the  general  experience 
of  steam  turbine  manufacturers  that  full  load  correction  curves, 
if  used  by  the  following  " ratio"  or  percentage  method,  can  be 
used  for  correcting  fractional  or  over  loads.  This  statement 
applies  at  least  without  appreciable  error  from  half  to  one  and  a 
half  load,  and  is  the  only  practicable  method  for  quarter  load  as 
well.*  Stated  in  a  few  words,  it  is  assumed  then  that  the  steam 
consumption  at  fractional  loads  is  changed  by  the  same  percent- 
age, as  at  full  load,  for  an  inch  of  vacuum,  a  degree  of  super- 
heat, or  a  pound  of  pressure.  It  will  now  be  shown  how  this 
method  applies  to  the  correction  of  the  steam  consumption  of 
the  turbine  at  fractional  loads.  Now  according  to  the  curve  in 
Fig.  75  the  steam  consumption  at  27.5  inches  (25.6  pounds)  must 

9  C   O 

obviously  be  multiplied  by  the  ratio  -*^-  ,f  of  which  the  numer- 

25.6 

ator  is  the  steam  consumption  at  28  inches  and  the  denominator 
at  27.5  inches,  to  get  the  equivalent  consumption  at  28  inches 
vacuum.  This  reasoning  establishes  the  proper  method  for 
making  corrections;  that  is,  that  the  base  for  the  percentage 
(denominator  of  the  fraction)  must  be  the  steam  consumption  at 
the  condition  to  which  the  correction  is  to  be  applied.  {  Similarly 
the  correction  ratio  to  change  the  consumption  at  50°  F.  .super- 

*}  C    O 

heat  to  o°  F.  is~-2i.,  and  to  correct  175  pounds  pressure  to  165 
24.0 

*  A  very  exhaustive  investigation  of  this  has  been  made  by  T.  Stevens  and  H.  M. 
Hobart  which  is  reported  in  Engineering,  March  2,  1906. 

f  Assuming  that  this  short  length  of  the  curve  may  be  taken  for  a  straight  line 
without  appreciable  error. 

%  In  nearly  all  books  touching  this  subject  so  important  to  the  practical, 
consulting,  or  sales  engineer,  the  alternative  method  of  taking  the  steam  consump- 
tion at  the  required  conditions  as  the  base  for  the  percentage  calculations  is 
implied.  By  such  a  method  percentage  correction  curves  derived  from  straight 
lines  like  Figs.  76  and  77  would  be  straight  lines  and,  in  application,  give  ab- 
surd results.  Actually  such  percentage  corrections  will  fall  on  curves  (see  Figs.  87 
and  88). 


METHOD   FOR   CORRECTING  STEAM  TURBINE  TESTS    131 


pounds  the  ratio  is-^1--     Data  and  calculated  results  obtained 

24-3 
by  this  method  may  then  be  tabulated  as  follows: 


Conditions 
of  Test. 

Required 
Conditions. 

Correction 
Ratio. 

Percentage 
Correction. 

Vacuum   inches   . 

27     ? 

28 

25.0 

—  2    34%* 

Superheat,  degrees  F  

SO. 

o 

25-6 
25.0 

+  4    17% 

Admission  pressure,  pounds  absolute  .  . 
Net  correction 

175- 

i65 

24.0 
24.8 
24-3 

+  2.06% 
+  -?   80% 

*  Steps  in  the  calculation  are  omitted  in  the  table,  thus  — '—  =  .0766  or  07.66  per  cent.,  making  the 

25.6 

correction  100  —  97.66,  or  2.34  per  cent.  It  may  seem  unreasonable  to  the  reader  that  these  percent- 
ages are  calculated  to  three  figures  when  the  third  figure  of  the  values  of  steam  consumption  is 
doubtful .  In  practice,  however,  the  ruling  of  the  curve  sheets  must  be  much  finer  and  to  larger  scale 
so  that  the  curves  can  be  read  more  accurately. 

The  signs  +  and  —  are  used  in  the  percentage  column  to 
indicate  whether  the  correction  will  increase  or  decrease  the 
steam  consumption.  "  Net  correction  "  is  the  algebraic  sum  of 
the  quantities  in  the  last  column. 

The  following  table  gives  the  results  of  applying  the  above 
"net  correction"  to  fractional  loads. 


£  Load 

\  Load 

|  Load 

|Load 

f  Load 

31-3  kw. 

62.5  kw. 

93.8  kw. 

i  25  kw. 

156.3  kw. 

Steam   consumption   from 

test  (Fig.  73)  

31.2 

26.9 

25.2 

24-5 

23-6 

Net  correction  +  3  .  89  %.  .  . 

+  1.2 

+  1.1 

+  1.0 

+  1.0 

+    .9 

Corrected  steam  consump- 

tion 

•32    A. 

28  o 

26    2 

2  C     ? 

24.    ^ 

Curve  B  in  Fig.  74  shows  the  corrected  curve  of  steam  con- 
sumption for  the  steam  turbine  as  plotted  from  the  above  table. 
By  thus  combining,  on  the  same  curve  sheet,  curves  A  and  B  as 
in  this  figure,  the  points  of  better  economy  of  the  turbine  are 
readily  understood. 


132 


THE    STEAM   TURBINE 


Results  of  economy  tests  of  the  various  turbines  given  on  the 
preceding  pages  are  of  very  little  value  for  comparison  when  the 
steam  consumptions  or  "  water  rates  "  are  given  for  all  sorts  of  con- 
ditions. With  the  assistance,  however,  of  curves  like  those  shown 
in  Figs.  75,  76,  and  77,  if  they  are  representative  of  the  type  and 
size  of  turbine  tested,  it  is  possible  to  make  valuable  compari- 
sons between  two  or  more  different  turbines.  Some  very  recent 
data  of  Curtis  and  Westinghouse-Parsons  turbines  are  given 
below,  together  with  suitable  corrections  adopted  by  the  manu- 
facturers for  similar  machines. 


Steam  Consumption 
Lbs.per  Kw.Hour 

t-L  t-i  |_L  |_k  M  M 

o  to  *-  os  oo  o 

x 

x 

X 

x^ 

^^ 

X 

X 

X. 

N, 

\ 

\ 

\ 

20212223342526272829 

Vacuum  inches  of  Mercury 

FIG.  78.    Typical  Vacuum  Correction  Curve  of  a  5ooo-Kilowatt  Impulse 

Turbine. 

The  following  test  of  a  Westinghouse-Parsons  turbine,  rated 
at  7500  kilowatts,  was  taken  at  Waterside  Station  No.  2  of  the 
New  York  Edison  Co.,  and  a  comparison  is  made  with  a  test 
of  a  five-stage  gooo-kilowatt  Curtis  turbine  at  the  Fisk  Street 
Station  of  the  Commonwealth  Electric  Company  of  Chicago. 
As  no  pressure  correction  is  given  for  the  Curtis  machine,  the 
New  York  Edison  test  is  corrected  to  the  pressure  at  which  the 
other  machine  was  operated  (179  pounds  per  square  inch  gauge). 
Approximately  an  average  vacuum  for  the  two  tests  is  taken 
for  the  standard,  and  100°  F.  superheat  is  used  for  comparing 
the  superheats.  These  assumed  standard  conditions  make  the 


METHOD   FOR   CORRECTING   STEAM  TURBINE  TESTS     133 

corrections  for  each  turbine  Comparatively  small.  When  two 
tests  are  to  be  compared,  by  far  the  more  intelligent  results  are 
obtained  if  each  is  corrected  to  the  average  conditions  of  the 
two  tests,  rather  than  correcting  one  test  to  the  conditions  of  the 
other.  There  is  always  a  chance  for  various  errors  when  large 
corrections  must  be  made. 


7500-KILOWATT    WESTINGHOUSE-P ARSONS    TURBINE,  WATERSIDE 
STATION  NO.  2,   NEW    YORK    EDISON    COMPANY. 
Tested  Sept.   i,    1907. 


Corrected 
to 

Correction, 
per  Cent.* 

8 

7CQ 

Average  steam  pressure,  pounds  gauge  
Average  vacuum,  inches  (referred  to  30  in.  barom.) 
Average  superheat   degrees  F 

177-5 
27-3 

nc   7 

179 
28-5 
IOO 

~     -15 
-3.36 
—        2O 

Average  load  on  generator   kilowatts 

Vo    / 
08  ?o   ? 

Steam  consumption   pounds  per  kilowatt-hour 

1C  .  ICT 

Net  correction   per  cent          

—  i.'io 

Corrected  steam  consumption,  pounds  per  kilo- 
watt-hour   

14.57 

*  The  following  corrections  were  given  by  the  manufacturers  and  accepted   by  the  purchaser  as 
representative  of  this  type  and  size  of  turbine: 
Pressure  correction  .,  per  cent,  for  x  pound. 
Vacuum  correction  3.8  per  cent  for  i  inch 
Superheat  correction  7.0  per  cent,  for  100  degrees  F. 
t  This  is  7$  per  cent,  better  than  the  manufacturer's  guarantee. 


November,  1907,  p.  658. 


9000-KILOWATT   CURTIS    TURBINE,    FISK    STREET   STATION,    COMMON- 
WEALTH   ELECTRIC    COMPANY,    CHICAGO.      Tested  in  1907. 


Corrected 
to 

Correction, 
per  Cent.* 

Speed  revolutions  per  minute  

7?o 

Average  steam  pressure,  pounds  gauge  
Average  vacuum,  inches  (referred  to  30  in.  barom.) 
Average  superheat  degrees  F 

179 

29-55 
116 

I79 

28.5 
IOO 

.O 
+    8.40 
+    1.28 

Average  load  on  generator  kilowatts 

8070. 

Steam  consumption   pounds  per  kilowatt-hour  . 

I^.O 

+  9  68 

Corrected  steam  consumption,  pounds  per  kilo- 
watt-hour   

14.26 

*  The  following  percentage  corrections  were  used : 
Superheat  correction  8  per  cent,  for  100°  F. 
Vacuum  correction  8  per  cent,  for  i  inch  from  curve  in  Fig. 
Pressure  correction  not  given . 


G.  E.  Bulletin, 

7O'      f  -VT 

No.  4531. 


134  THE   STEAM  TURBINE 

These  results  show  a  difference  of  only  .20  pound  in  the  cor- 
rected steam  comsumption,  so  that  for  exactly  the  same  con- 
ditions these  two  machines  would  probably  give  approximately 
the  same  economy.  Each  turbine  is  doubtless  best  for  the 
special  conditions  for  which  it  was  designed. 

These  results  are  equivalent  to  respectively  9.58  pounds  and  9.72 
pounds  per  indicated  horsepower,  assuming  97  per  cent  as  the  effi- 
ciency of  the  generator  and  91  per  cent  as  the  mechanical  efficiency 
of  a  large  Corliss  engine  according  to  figures  given  by  Stott.* 

From  experience  with  other  similar  turbines  it  seems  as  if  the 
vacuum  corrections  given  are  too  low  for  each  turbine.  The 
correction  for  the  Curtis  turbine  was  obtained  from  the  curve  in 
Fig.  78  as  given  between  27  and  28  inches,  while  it  was  used 
between  28.5  and  29.5  inches,  where  the  curve  of  steam  consump- 
tion most  likely  slopes  somewhat  as  shown  by  the  dotted  curve  in 
the  figure,  which  was  derived  from  the  percentage  change  of 
the  theoretical  steam  consumption  calculated  from  the  available 
energy.  The  correction  of  2.7  per  cent,  per  inch  of  vacuum 
for  the  Westinghouse-Parsons  turbine  is  probably  too  low  also, 
although  the  percentage  correction  would  not  be  nearly  as 
large  as  for  the  Curtis.  If  both  of  these  corrections  are  too 
low,  the  effect  of  increasing  them  would  be  to  increase  the  cor- 
rected steam  consumption  of  the  Curtis  turbine  and  reduce  that 
of  the  Westinghouse-Parsons. 

Large  sizes  of  steam  turbines  are  also  made  by  the  Allis- 
Chalmers  Company,  but  sufficient  data  are  not  given  with  pub- 
lished tests  to  make  a  comparison  here. 

Tests  of  a  5ooo-kilowatt  Curtis  and  a  75ookilowatt  Westing- 
house-Parsons turbine  are  also  recorded  here  for  comparison. 
The  two  tests  are  corrected  to  the  assumed  standard  conditions 
of  173.7  pounds  gauge  pressure,  28  inches  vacuum,  and  o°  F. 
superheat.  For  the  test  of  the  Curtis  machine  the  same  per- 
centage corrections  were  used  as  for  the  9Oookilowatt  turbine; 

*  Electric  Journal,  July,  1907.  It  is  stated  also  in  this  article  that  the  vacuum 
correction  of  a  Westinghouse-Parsons  turbine  is  3.5  per  cent,  per  inch  betv/een  28 
and  28.5  inches.  Jude  states  that  the  vacuum  correction  for  Parsons  turbines  is 
five  to  six  per  cent. 


METHOD  FOR  CORRECTING  STEAM  TURBINE  TESTS    135 


and  for  the  test  of  the  Westinghouse  turbine  the  vacuum  correc- 
tion is  that  given  in  the  footnote  at  the  bottom  of  page  134  (3.5 
per  cent,  per  inch),  while  the  other  percentage  corrections  are  the 
same  as  in  the  preceding  test  of  a  similar  machine.  The  West- 
inghouse turbine  was  operated  with  wet  steam.  In  a  test  of  a 
reciprocating  engine  the  equivalent  economy  with  dry  steam  is 
calculated  by  merely  subtracting  the  percentage  of  moisture,  but 
in  a  turbine  test  the  correction  is  generally  stated  as  being  a  little 
more  than  twice  the  percentage  of  moisture.  In  other  words,  in  a 
turbine  test  the  moisture  must  be  subtracted  twice.  The  reason 
for  this  difference  in  the  methods  of  correcting  water  rates  of 
engines  and  turbines  is  the  very  large  increase  in  the  disk  and 
blade  rotation  losses  in  wet  steam  (cf.  Fig.  69). 

5000-KILOWATT     FIVE-STAGE     CURTIS     TURBINE,     L    STREET     STATION, 
BOSTON    EDISON    COMPANY.     Tested  Jan.  29,  1907. 


1 

Corrected 
to 

Correction, 
per  Cent. 

Duration  of  test   hours 

Speed   revolutions  per  minute 

720 

177.  7 

177.  7 

O 

Average  vacuum,  in.  (referred  to  30  in.  barom.)  .  .  . 
Average  superheat   degrees  F     

28.8 

142 

28 
o 

+    6.40 
+  11.36 

Average  load  on  generators  kilowatts    

CIQC 

Steam  consumption   pounds  per  kilowatt-hour  .  . 

I?.  C2 

+  17.76 

Corrected  steam  consumption,  pounds  per  kilo- 

I  ^    02 

7500-KILOWATT     WESTINGHOUSE-PARSONS      TURBINE      (SINGLE      FLOW 

TYPE),    INTERBOROUGH    RAPID    TRANSIT    COMPANY,    NEW    YORK. 

Tested  in  1907. 


Corrected 
to 

Correction, 
per  Cent. 

3 

Average  steam  pressure   pounds  gauge 

149-7 
27.70 

3-o 
7135 

17.  70 

173-7 
28 
0 

—  24. 

Average  vacuum,  in.  (referred  to  30  in.  barom.)  .  . 

-1.05 

—  6.0 

Steam  consumption,  pounds  per  kilowatt-hour 

-9-45 

Corrected  steam  consumption,  pounds  per  kilo- 

16.  10 

136 


THE  STEAM  TURBINE 


It  is  stated  that  the  steam  consumption  of  the  Interborough 
Company's  turbine  is  15.87  pounds  at  full  load  and  15.54  pounds 
at  9000  kilowatts  when  the  overload  valve  opens.  The  gen- 
erator connected  to  this  turbine  is  rated  at  only  5500  kilowatts. 
With  a  generator  more  nearly  the  rating  of  the  turbine  it  is 
probable  still  better  results  would  be  secured. 

Corrected  tests  of  a  20oo-kilowatt  Curtis  and  a  looo-kilowatt 
Westinghouse-Parsons  turbine-generator  are  also  given  here. 
Assumed  standard  conditions  and  corrections  are  taken  the  same 
as  in  the  two  tests  preceding,  except  that  the  Westinghouse  test 
is  corrected  to  "the  steam  pressure  of  the  Curtis  test. 

2000-KILOWATT  CURTIS  TURBINE,  COMMONWEALTH  ELECTRIC  COMPANY 
CHICAGO.      Tested  May,   1905,  by  Sargent  &  Lundy. 


Corrected 
to 

Correction, 
per  Cent. 

Duration  of  test    hours  

I     2s 

Speed,  revolutions  per  minute    

ooo 

Average  steam  pressure,  pounds  gauge  
Average  vacuum,  in.  (referred  to  30  in.  barom.)  .  . 
Average  superheat   degrees  F 

166.3 

28.5 

2O7 

166.3 
28 
o 

o 
+    4.0 
4-  16  s? 

Average  load  on  generators   kilowatts 

2O24 

Steam  consumption,  pounds  per  kilowatt-hour 

I  S    O2 

Net  correction  

4-  20    ?  ~l 

Corrected  steam  consumption,  pounds  per  kilo- 
watt-hour   

18.10 

1000- KILO  WATT     WESTINGHOUSE-PARSONS     TURBINE. 

1907,   by  S.   Gilliard. 


Tested     September, 


Corrected 
to 

Correction, 
per  Cent. 

Duration  of  test   hours     

i 

Speed,  revolutions  per  minute    

1800 

Average  steam  pressure,  'pounds  gauge  

147.  6 

166.? 

-I    87 

Average  vacuum,  in.  (referred  to  30  in.  barom.)..  . 
\verage  moisture    per  cent          .            ... 

27.02 

7? 

28 
o 

-3-40 
—  i   co 

Average  load  on  water  brake,  horsepower  

ICO3.  C 

Equivalent  average  load  in  kilowatts  (generator 
efficiency  QJ.%)) 

IOs  < 

Steam   consumption,  pounds   per  brake   horse- 
power-hour (wet)  

1  3  61 

Steam  consumption,  pounds  per  equivalent  kilo- 
watt-hour (wet) 

IO    3? 

-6.77 

Corrected  steam  consumption,  pounds  per  kilo- 
watt-hour   

18.04 

METHOD  FOR  CORRECTING   STEAM   TURBINE  TESTS      137 

Curves  in  Fig.  79  are  given  to  compare  the  steam  consump- 
tion of  a  standard  5ooo-kilowatt  turbine-generator  and  a  4-cyl- 
inder  compound  5ooo-kilowatt  reciprocating  steam  engine  of 


\ 


«     O 
•£     of 

it 

I? 
Is" 

i.8 


51 


8- 


the  type  used  by  the  Interurban  and  Metropolitan  Companies 
of  New  York,  assuming  both  units  operating  under  the  same 
conditions.  These  curves  illustrate  the  good  overload  economy 


THE  STEAM  TURBINE 


of  the  turbine,  showing  that  at  50  per  cent,  overload  the  engine 
designed  for  equal  work  in  the  cylinders  requires  for  the  same 
output  43  per  cent,  more  steam  than  the  turbine. 

These  results  are  particularly  interesting  because  the  peak 
capacity  of  a  station  with  a  given  equipment  of  boilers  and 
auxiliaries  is  increased  in  proportion  to  the  reduction  of  steam 
consumption  at  overloads. 

For  a  given  investment  the  turbine  gives  a  much  larger  range 


800        1200 
Speed -H.P.M. 


2000       2400 


FIG.  80.    Torque,  Speed  Output,  and  Efficiency  Curves  of  a  Typical 
5oo-Kilowatt  Steam  Turbine. 

of  load  and,  moreover,  affords  the  means  by  which  the  peak 
capacity  of  existing  stations  can  be  greatly  increased. 

The  speed  output  curve  (Fig.  80)  is  very  useful  to  engineers  to 
determine  if  a  turbine  is  running  at  its  best  speed.  If  the  cor- 
responding curves  of  steam  consumption  per  kilowatt  output 
(usually  called  water  rate  per  kilowatt)  and  efficiency  are  calcu- 
lated according  to  the  form  on  page  272,  a  great  deal  of  informa- 
tion is  obtained  about  the  operation  and  economy  of  a  turbine. 
The  torque  line  in  Fig.  80  is  always  drawn  straight,  just  as  a 


METHOD  FOR  CORRECTING  STEAM  TURBINE  TESTS     139 

Willans  flow  line.  A  curve  of  total  steam  consumption  is  usually 
a  straight  line  for  the  normal  operating  limits  of  a  turbine,  but 
usually  becomes  curved  when  a  by-pass  valve  opens  on  overload, 
or  when  the  turbine  is  over  its  capacity  so  that  the  pressures 
are  not  normal  in  the  stages. 

The  torque  line  shows  why  a  turbine  engine  is  not  adaptable  to 
automobiles.  The  starting  torque  of  a  small  commercial  tur- 
bine is  not  large,  so  that  starting  would  be  difficult  with  a  small 
wheel,  and  reversing  and  speed  reduction  would  be  as  difficult 
as  with  a  gasoline  engine.  The  reciprocating  steam  engine  as 
well  as  the  gasoline  engine  has,  therefore,  advantages  over  the 
steam  turbine  for  this  service. 


CHAPTER  VII. 
COMMERCIAL  TYPES. 

IN  some  respects  the  order  in  which  the  commercial  types  of 
steam  turbines  are  discussed  on  the  following  pages  is  somewhat 
arbitrary;  but,  essentially,  it  is  in  the  order  of  relative  simplicity. 
De  Laval  and  Parsons,  of  the  modern  designers,  were  first  in 
the  field.  They  were  in  fact  pioneers  in  the  development  of 
commercial  steam  turbines,  and  other  designers  have*  followed 
more  or  less  in  their  steps.  The  reasons  for  giving  precedence 
to  the  types  which  they  developed  are  therefore  obvious,  and  no 
other  explanation  is  needed. 

Because  of  its  greater  simplicity  the  commercial  De  Laval  is 
first  discussed,  and  is  followed  with  descriptions  of  the  various 
forms  of  the  Parsons  turbine  and  the  more  recent  types. 

DE  LAVAL  STEAM  TURBINE. 

Rational  engineering  development  is  nowhere  better  exempli- 
fied than  in  the  successful  performance  of  the  De  Laval  steam 
turbine.  In  nearly  every  respect,  even  to  details,  it  is  still  prac- 
tically the  same  as  the  turbines  designed  under  the  personal 
direction  of  De  Laval. 

The  essential  elements  of  this  turbine  are:  (i)  the  nozzles  in 
which  the  steam  expands;  (2)  awheel  or  disk  with  suitable  blades 
on  its  periphery;  (3 )  a  slender  shaft  on  which  the  wheel  is  mounted; 
and  (4)  a  set  of  reducing  gears  to  change  the  high  speed  of  the 
turbine  shaft  to  a  lower  speed  adaptable  for  driving  machinery. 

Drawings  of  a  small  De  Laval  turbine  are  shown  in  Fig.  82. 
The  turbine  wheel,  W,  is  supported  upon  the  flexible  shaft 
between  the  bearing,  Z,  provided  with  a  spherical  seat,  and  the 
gland  or  stuffing-box,  P.  Teeth  are  cut  into  the  metal  of  the 
turbine  shaft  to  make  the  pinions  on  each  side  of  K  fit  the  gear 

140 


COMMERCIAL  TYPES 


141 


wheels  A  and  B,  from  which  the  power  is  transmitted.  The 
design  shown  here  is  intended  for  driving  two  electric  generators 
which  are  direct-connected  by  means  of  the  couplings  shown  at 
the  left  in  the  figure. 

De  Laval  turbine-generator  sets  of  from  50  horsepower  up- 
wards are  supplied  with  two  gear  wheels,  two  power  shafts,  and 
two  dynamos  for  each  turbine  wheel,  while  the  smaller  sizes 


o«,veN  T 

COUPLING      C  / 

I  'omviNtt 

>  COUPLING 


FIG.  82.    Section  of  a  De  Laval  Turbine  with  Two  Power  Shafts 
(on  Gear  Wheels  A  and  B). 


have  gear  arrangements  for  a  single  generator.  Because  of  the 
higher  speed  at  which  the  small  sizes  operate  (see  page  144), 
making  the  pressure  on  the  gear  teeth  considerably  smaller  than 
with  the  larger  sizes,  more  power  can  be  transmitted  with  a  single 
set  of  gears.  The  large  size  of  the  gear  wheels  compared  with 
the  turbine  is  a  noticeable  feature  of  these  turbines. 

Turbine  Wheel.  On  account  of  the  very  high  speeds  at  which 
these  turbines  operate,  the  wheels  or  disks  require  very  careful 
designing.  In  the  small  and  medium  sizes,  a  wheel  similar  to 


142 


THE  STEAM  TURBINE 


the  drawing  in  Fig.  83  is  used.  When  this  design  is  used,  the 
hub  of  the  wheel  is  bored  out  and  a  thin  steel  bushing  is  drawn 
into  it  by  means  of  a  nut  shown  in  the  figure  at  the  right-hand  end. 
Before  this  bushing  is  put  into  the  wheel,  it  is  forced  on  the  shaft 


FIG.  83 


De  Laval  Turbine  Wheel  with  a  Hole  at  the  Center  and  Details  of  the 
Blades. 


and  pinned  in  place  as  shown.     The  wheel  can  be  removed  from 

the  shaft  by  taking  off  the  nut  and  drawing  it  from  the  bushing. 

The  strength  of  a  disk,  or  a  wheel  of  a  disk  type,  in  which  there 

is  a  hole  at  the  center  is  at  best  not  more  than  half  as  strong  as  one 


FlG.  84.    De  Laval  Turbine  Wheel  without  a  Hole  at  the  Center. 

without  a  hole.*  On  this  account  in  the  larger  sizes  of  De  Laval 
turbines  it  has  been  found  necessary  to  use  the  design  shown 
in  Fig.  84.  In  this  arrangement  a  solid  disk  is  permitted. 
The  hub  is  recessed  at  each  end,  and  the  flexible  shaft  is  made 

*  An  explanation  of  this  remarkable  phenomenon  is  given  on  page  333. 


COMMERCIAL  TYPES  143 

with  enlarged  flanged  ends  which  fit  into  the  recesses  and  are 
bolted  solidly  in  place.  The  recesses  and  flanges  are  machined 
with  a  four  per  cent,  taper  in  order  that  the  parts  maybe  accurately 
centered  and  fitted. 

This  form  of  wheel  disk  with  the  section  increasing  from  the 
rim  towards  the  hub  is  arrived  at  by  proportioning  it  to  have 
equal  unit  stresses  throughout.  But  this  condition  does  not 
hold  true  at  the  rim,  where  just  below  the  blades  annular 
grooves  are  turned  on  each  side.  Weakening  of  the  wheel  at 
the  rim  is  a  very  good  method  of  providing  for  abnormal  stresses 
that  result  in  case  of  a  failure  of  the  governor  to  control  the  speed. 
The  purpose  in  making  these  grooves  is  to  have  the  wheel  burst 
at  this  reduced  section  where  the  stresses  per  unit  of  area  are 
about  50  per  cent,  larger  than  at  any  other  part  of  the  wheel, 
rather  than  near  the  center  where  the  damage  from  failure  would 
be  so  much  greater.  At  normal  speed  the  factor  of  safety,  at 
this  smallest  section,  is  about  five,  and  since  the  unit  stresses 
vary  as  the  square  of  the  speed,*  the  wheel  will  fail  at  this  place 
at  a  little  more  than  twice  the  rated  speed.  As  these  wheels  are 
constructed,  no  great  damage  to  the  turbine  will  result,  therefore, 
from  the  failure  of  the  wheel  rim.  It  has  been  shown  by  actual 
experiments  with  such  wheels  that  when  failure  occurs,  the  rim 
holding  the  blades  is  broken  up  into  very  small  pieces  which 
will  not  damage  the  wheel  case.  It  is  stated,  however,  that 
wheels  without  this  reduced  section,  when  tested  to  failure,  have 
been  broken  up  into  two  or  three  large  pieces  by  bursting  through 
the  center,  and  these  pieces  have  been  driven  through  an  experi- 
mental wheel  casing  made  of  two-inch  steel  castings. 

There  is  also  another  consideration  that  is  especially  interesting 
to  engineers.  When  a  portion  of  the  rim  breaks  off  the  wheel 
becomes  unbalanced,  and  as  the  clearance  between  the  heavy 
hub  of  the  wheel  and  the  safety  bearings  in  the  surrounding 
casing  is  very  small,  as  can  be  seen  in  Fig.  82,  the  flexibility  of  the 

IFF2 
*  Centrifugal  force  =  —    -   (see  page  314)  and  is  therefore  proportional  to  the 

square  of  velocity  (speed).  The  factor  of  safety  at  other  sections  of  a  De  Laval 
wheel  is  about  eight. 


144 


THE  STEAM  TURBINE 


shaft  will  permit  the  hub  of  the  wheel  to  come  into  contact  with 
the  circular  openings  in  the  casing  into  which  it  extends.  The 
friction  of  these  surfaces  will  act  as  a  brake  and  assists  in  bringing 
the  wheel  to  rest.  And  this  is  easily  accomplished,  because  with 
the  blades  removed  the  steam  no  longer  acts  to  rotate  the  wheel. 
The  diameters  of  the  wheels  are  relatively  small,  as  can  be  seen 
from  the  following  table: 


Horsepower             .                

5" 

30 

IOO 

?oo 

Revolutions  per  minute  

30,000 

20,000  « 

13,000 

IO  OOO 

Diameter  to  center  of  blades,  inches.  .  . 

3-94 

8.86 

19.68 

29.92 

Wheels  for  De  Laval  turbines  are  usually  made  of  a  special 
forged  nickel  steel  said  to  be  rather  high  in  carbon. 

Nozzles.  Fig.  85  is  a  typical  illustration  of  a  20-kilowatt 
De  Laval  turbine-generator  and  gives  a  general  idea  of  how  the 
nozzles  which  direct  the  steam  against  the  blades  are  arranged 
around  the  periphery  of  the  turbine  wheel.  They  are  attached 
to  the  turbine  mechanically  by  being  fitted  into  the  circumfer- 
ence of  the  steel  casting  which  serves  as  the  casing  for  the  wheel. 
The  number  of  nozzles  varies  according  to  the  size  of  the  turbine. 
The  nozzles  are  provided  with  hand  valves,  which  can  be  seen 
in  the  figure,  by  which  they  can  be  closed  when  the  turbine  is 
running  at  light  loads.  In  this  way  some  of  the  nozzles  are  "cut 
out"  and  a  relatively  high  efficiency  is  obtained  at  light  loads. 
In  this  particular  case,  about  half  of  the  openings  in  the  casing 
for  nozzles  are  closed  by  plugs;  but  by  removing  these  plugs 
and  inserting  nozzles  instead,  the  capacity  of  the  turbine  would 
be  greatly  increased. 

The  nozzles  are  the  only  parts  of  a  De  Laval  turbine  that 
are  changed  to  make  it  suitable  for  any  particular  pressure, 
degree  of  superheat,  or  vacuum.  The  ratio  of  the  admission 
(usually  boiler)  pressure  to  the  exhaust  pressure  is  the  most  im- 
portant factor  influencing  the  design  of  a  nozzle.  Briefly  stated 
this  ratio  of  pressures  determines  the  areas  of  the  cross-section 
of  the  nozzle  at  the  throat  and  at  the  mouth,  and  therefore  its 
divergence  or  taper. 


COMMERCIAL  TYPES 


145 


For  the  same  output  more  steam  is  required  at  a  low  pressure 
than  at  a  higher  pressure.  De  Laval  turbines  are  readily  adjusted 
for  a  change  of  boiler  pressure  by  adding  more  nozzles  if  they  are 


needed.  Sometimes  turbines  are  fitted  with  two  sets  of  nozzles, 
one  suitable  for  condensing  and  the  other  for  non-condensing 
operation. 


146  THE  STEAM  TURBINE 

Reamers  are  used  to  produce  the  required  taper  on  the  inside 
of  these  nozzles.  In  the  works  at  Trenton  over  600  reamers  are 
kept  in  the  tool  room.  The  taper  of  the  nozzle  ranges  from  six 
to  twelve  degrees,  and  the  clearance  between  the  mouth  of  the 
nozzle  and  the  blades  (axial  clearance)  is  about  an  eighth  of  an 
inch. 

Blades.  De  Laval  blades  are  made  of  drop-forged  steel  and 
have  bulb  shanks  which  are  fitted  into  suitable  slots  in  the  wheel, 
shown  in  Fig.  83,  which  are  milled  across  the  rim  and  then  drilled. 
The  blades  are  lightly  calked  to  secure  them  in  place.  At  the 
upper  ends  of  the  blades  they  are  provided  with  "extensions" 
which  are  designed  to  make  adjoining  blades  fit  closely  and  thus 
form  a  continuous  ring  over  the  blades  at  the  periphery  of  the 
wheel.  Details  of  these  blades  are  shown  more  clearly  in 
Fig.  64. 

Shaft.  De  Laval  steam  turbines  have  two  important  features 
distinguishing  them  from  all  other  types.  The  first  is  the  long 
diverging  nozzle  with  the  hand  wheel  control  already  mentioned; 
and  the  second  is  the  slender  flexible  shaft  *  of  the  turbine.  A 
wheel  revolving  at  a  very  high  speed  tends  to  rotate  about  its 
center  of  gravity.  If  it  is  mounted  on  a  stiff,  unyielding  shaft,  of 
which  the  axis  does  not  pass  through  the  center  of  gravity  of  the 
wheel,  this  tendency  causes  violent  vibrations  of  the  wheel  and 
shaft  due  to  the  very  large  centrifugal  forces.  It  is  stated  that  a 
weight  of  one  ounce  attached  at  the  circumference  of  the  wheel  of 
a  3oo-horsepower  De  Laval  turbine  will  produce  an  unbalanced 
centrifugal  force  of  nearly  2000  pounds.  It  is  mechanically 
difficult  and  almost  impossible  to  construct  a  wheel  so  perfectly 
balanced  that  its  center  of  gravity  will  exactly  coincide  with 
the  geometric  center  of  the  shaft  on  which  it  is  mounted. 
De  Laval,  therefore,  devised  a  long,  slender  shaft  which,  as  the 
speed  of  the  wheel  increases,  yields  somewhat  and  allows  the 
latter  to  assume  its  own  position  of  rotation  about  its  center  of 
gravity. 

*  The  diameter  of  the  shaft  of  a  loo-horsepower  De  Laval  turbine  is  i  inch  and 
of  a  30ohorsepower  turbine  is  about  i^  inches. 


COMMERCIAL  TYPES  147 

The  wheel  is  not  mounted  midway  between  the  bearings  but 
considerably  nearer  the  spherical  seated  bearing  Z,  Fig.  82,  at 
the  governor  end.  Now  when  the  wheel  is  started  up  from  rest, 
if  its  center  of  gravity  is  not  precisely  in  the  axis  of  the  shaft,  it 
will  bend,  and  the  plane  of  revolution  of  the  wheel  is  then  no 
longer  perpendicular  to  the  axis  of  rotation.  When,  however,  a 
sufficiently  high  speed  is  reached,  so  that  gyroscopic  action  is 
great  enough  to  pull  this  plane  back  to  a  position  perpendicular 
to  the  axis  of  rotation,  a  "node"  is  formed  at  the  center  of  the 
hub  and  rotation  will  then  take  place  about  the  center  of  gravity 
of  the  system.  The  speed  at  which  the  amplitude  of  vibration 
is  greatest  is  called  critical.* 

Bearings.  Typical  bearings  of  De  Laval  turbines  are  illus- 
trated in  the  section  drawings  in  Fig.  82.  At  the  right-hand  or 
"governor"  end  there  is  a  spherical  seated  bearing  (Z).  A 
design  of  this  kind  is  used  for  the  purpose,  primarily,  of  giving 
greater  flexibility  to  the  shaft  and  to  take  the  small  end  thrust 
exerted  on  the  wheel  by  the  steam  issuing  from  the  nozzles  at  a 
very  high  velocity.  In  single  wheel  turbines  of  the  De  Laval 
type  this  pressure  or  thrust  is,  however,  very  slight,  as  the  steam 
is  expanded  to  the  exhaust  pressure  before  it  leaves  the  nozzles. 
It  is  obvious,  therefore,  that  the  wheel  rotates  in  steam  of  very 
nearly  the  same  pressure  on  both  of  its  sides.  Such  a  design 
has  also  the  advantage  of  being  self-aligning.  A  helical  spring 
shown  in  the  same  figure  holds  the  spherical  bearing  against  its 
seat  in  the  turbine  casing.  On  the  other  side  of  the  turbine 
wheel  the  shaft  passes  through  a  loose-fitting  bearing,  P,  serving 
primarily  as  a  gland  or  stuffing-box  to  prevent  the  leakage  of 
steam  from  the  casing.  The  shaft  does  not  pass  through  the 
casing  on  the  right-hand  side,  so  that  no  precautions  are  necessary 
to  prevent  leakage  of  steam  on  that  side.  At  each  side  of  the 
pinions  of  the  reduction  gearing,  the  turbine  shaft  is  supported 
on  plain  white-metal  (Babbitt)  bearings  C  and  CC.  The  sur- 

*  "Critical  speed"  is  the  name  given  to  that  speed  of  a  wheel  at  which  it  tends 
to  rotate  about  its  own  center  of  gravity.  In  the  De  Laval  turbines  it  occurs  at 
about  ^  to  |  of  the  normal  running  speed. 


148 


THE   STEAM  TURBINE 


face  speed  in  these  bearings   is   usually  designed   to   be   about 
70  feet  per  second. 


• 

bO 

3 

^    'O 

l-l  (U 

c    o 

O        ?H 

^    I 
I  I 

^a 


i 
i 

o 


Speed-reduction  Gears.  On  account  of  the  high  speed  of  the 
turbine  shaft,  reduction  gears  are  required  to  bring  the  speed 
within  practicable  limits  for  utilizing  the  power.  The  reduction 


COMMERCIAL  TYPES  149 

is  usually  about  ten  to  one,  anji  is  accomplished  by  means  of 
small  pinions  on  the  turbine  shaft  meshing  with  steel  helical 
gear  wheels.  The  teeth  of  the  pinions  are  very  small  and  are 
cut  directly  into  an  enlarged  section  of  the  flexible  shaft.*  The 
teeth  for  this  gearing  are  cut  spirally  at  an  angle  of  45  degrees. 
As  indicated  in  Fig.  86  the  teeth  on  one  side  are  cut  on  a  right- 
hand  and  on  the  other  side  on  a  left-hand  spiral.  This  method 
effectually  prevents  any  movement  of  the  shaft  in  the  direction 
of  the  axis  and  balances  the  thrust  of  the  gears.  Previous  to 
the  time  when  De  Laval  demonstrated  that  gears  could  be  oper- 
ated at  a  linear  velocity  of  more  than  100  feet  per  second,  the 
high  speeds  which  he  introduced  were  not  considered  practically 
possible.  His  success. at  these  high  speeds  was  due  largely  to 
the  fine  pitch  f  and  spiral  angle  of  the  teeth.  It  is  thus  possible 
to  bring  a  large  number  of  teeth  into  mesh  at  the  same  time, 
so  that  the  working  pressure  on  each  tooth  is  made  very  small 
and  abrasion  is  reduced  to  a  minimum. 

The  reduction  gears  are  enclosed  in  a  casing  entirely  separate 
from  that  around  the  turbine  wheel.  This  casing  prevents  dust 
and  grit  from  getting  into  the  gears  and  avoids  accidents  from 
persons  or  objects  falling  upon  them.  With  careful  attention 
these  gears  sometimes  run  for  several  years  without  visible  wear. 
Formerly  the  gear  wheels  were  made  of  bronze,  but  experience 
showed  that  the  teeth  became  crystallized  after  a  few  years  of 
operation,  and  pieces  of  the  teeth  which  were  sometimes  broken 
off,  were  liable  to  injure  other  teeth.  Such  gears  should  always 
be  supplied  with  a  little  oil  for  lubrication. 

This  speed-reduction  gearing  introduces  two  important  dis- 
advantages: first,  the  friction  loss  is  considerable;  and  second, 
the  construction  is  necessarily  expensive.  The  friction  loss, 
obviously,  will  depend  largely  on  the  quality  of  workman- 
ship. It  is  stated  that  this  loss  in  the  gears  is  about  5  per 

*  The  pinions  are  said  to  be  made  of  .60  to  .70  carbon  steel,  and  the  teeth  of  the 
larger  gear  wheels  are  cut  in  .20  carbon  steel  of  a  grade  similar  to  that  used  for 
locomotive  wheel  tires. 

t  The  pitch  of  the  gears  varies  from  .15  inch  in  the  smallest,  to  .26  inch  in  the 
largest  sizes. 


ISO 


THE  STEAM  TURBINE 


cent.*  of  the  power  transmitted  when  they  are  in  good  con- 
dition, and  sometimes  as  much  as  10  per  cent,  in  moderately 
worn  gears. 

After  a  few  years  of  service  it  is  usually  found  that  the  steam 
consumption  of  a  De  Laval  turbine  is  slightly  greater  than  when 
it  was  new.  This  poorer  economy  is  probably  due  to  the  increased 
loss  in  the  gears  from  wear  as  well  as  to  the  wearing  away  of  the 


.30 


20        40 


80       100      120      140 
Superheat-Degs.  Fahr. 


160       180       200 


FIG.  87.     Percentage  Curve  for  Correcting  De  Lava!  Turbine  Tests  for 

Superheat. 


blades  on  the  turbine  wheel,  which  by  changing  the  shape  of  the 
blades  causes  a  loss  of  efficiency. 

Governor.  The  De  Laval  governor  is  shown  in  Figs.  162 
and  163,  page  220,  where  methods  of  governing  are  discussed. 
The  valve  arrangement  controlled  by  the  governor  is  a  plain 
throttling  type. 

*  Regarding  these  losses  the  results  of  experimenters  differ  a  great  deal.  Lewicki 
found  the  gearing  and  bearing  loss  in  a  3o-horsepower  De  Laval  turbine-generator 
to  be  7.5  per  cent,  of  the  full  load  output.  Delaporte  states  that  the  gearing  losses 
of  a  200-horsepower  De  Laval  turbine  are  about  i  per  cent,  when  new;  and  he  states 
also  that  in  his  opinion  the  combined  gearing  and  bearing  friction  losses  of  a  300- 
horsepower  De  Laval  turbine  should  be  taken  roughly  at  about  3  per  cent. 


COMMERCIAL   TYPES  !$! 

Superheat,  Vacuum,  and  Economy  Curves.  Fig.  87  shows  by 
percentages  the  effect  of  superheat  on  the  steam  consumption. 
For  low  values  of  superheat  the  gain  for  a  De  Laval  turbine  is 
much  greater  than  for  larger  amounts  of  superheat.  Such  curves  on 
a  percentage  basis  are  sometimes  very  serviceable  to  show  strik- 
ing variations  clearly.  Fig.  88  is  a  similar  percentage  curve  to 
show  how  the  vacuum  influences  the  steam  consumption.  With 
a  high  vacuum  the  improvement  in  economy  is  much  more 


10        12       14       16       18 
Vacuum  in  Inches  of  Mercury 


24       26 


FIG.  88.     Percentage  Curve  for  Correcting  De  Laval  Turbine  Tests  for 

Vacuum. 


marked  than  at  low  values.  Fig.  89  shows  approximately  the 
steam  consumption  for  any  size  of  De  Laval  turbine-generator 
operating  non-condensing  or  with  28  inches  vacuum  at  165  pounds 
per  square  inch  absolute  pressure,  and  o  degrees  F.  superheat. 

It  is  stated  that  the  half  load  steam  consumption  of  a  De  Laval 
turbine  is  12  percent,  greater  than  the  full  load  value,  and  that 
at  quarter  load  it  is  25  per  cent,  more  than  that  at  full  load. 
For  such  good  performance  at  light  loads  it  is  necessary  to 
operate  the  turbine  with  no  more  valves  open  than  are  needed. 


THE  STEAM-TURBINE 


Because  the  valves  must'  be  operated  by  hand  such  good  economy 
could  probably  not  be  obtained  with  a  rapidly  fluctuating  load. 


30 


10 


GO 


80         100        120        140        160        180        300        230        240 
Hated  Full  Load  Output— Kw. 

FIG.  89.    Approximate  Steam  Consumption  of  any  Size  of  De  Laval 

Turbine-generators.     Dry  Saturated  Steam  at  165  Pounds 

pej  Square  Inch  Absolute  Pressure. 

Turbine  Losses.  The  following  table  shows  how  the  losses  in 
a  De  Laval  2oo-kilowatt  turbine-generator  have  been  divided  up 
by  Stevens  and  Hobart: 

Nozzle  losses 12  per  cent. 

Radiation  losses  and  leakage i    "      " 

Rotation  losses  due  to  the  turbine  wheel  revolving  in  steam 

Losses  due  to  the  steam  traveling  over  the  blades 

Bearing  friction  losses 

Losses  in  speed-reduction  gearing 

Generator  losses 

Losses  due  to  residual  kinetic  energy  in  the  steam  passing 

to  the  condenser 8 

Electrical  output 59 

Total..  ioo 


COMMERCIAL  TYPES  153 

PARSONS'  TURBINE 

The  Parsons  type  of  steam  turbine  differs  from  that  commonly 
known  as  De  Laval's  principally  in  the  substitution  of  stationary 
blades  in  the  place  of  nozzles.  These  stationary  blades  are  so 
shaped  as  to  direct  the  steam  upon  the  moving  blades  just  as 
nozzles  would.  In  turbines  of  this  type  a  large  number  of  rows 
of  moving  blades  are  employed,  which  are  attached  to  the  cylin- 
drical surface  of  a  revolving  drum,  called  a  rotor. 

There  is  also  another  difference  which,  from  a  theoretical  view- 
point, makes  a  Parsons  turbine  entirely  different  from  other  types. 
All  the  impulse  turbines,  of  which  the  De  Laval  is  a  good  example, 
make  very  little,  if  any,  provision  for  the  expansion  of  the  steam 
in  the  moving  blades,  while  the  Parsons  type  is  designed  to  give 
approximately  as  much  expansion  of  the  steam  in  the  moving 
as  in  the  stationary  or  " guide"  blades.  In  turbines  of  this 
type  each  set  of  one  row  of  moving  and  one  row  of  stationary 
blades  is  called,  technically,  a  stage. 

Compared  with  the  De  Laval  turbine  in  which  the  blades  of 
a  single  wheel  revolve  in  a  medium  of  uniformly  low  density 
with  the  pressure  very  nearly  the  same  on  both  sides  of  the  wheel, 
most  of  the  blades  of  a  Parsons  turbine  revolve  in  steam  of  high 
density.  Blades  at  the  admission  end  revolve  in  steam  at  very 
nearly  the  boiler  pressure,  and  only  those  at  the  low-pressure  end' 
are  in  steam  of  low  density. 

In  the  Parsons  turbine,  because  of  the  large  number  of  blades, 
many  of  which  revolve  in  steam  of  very  high  density,  the  disk 
and  blade  rotation  losses  are  very  much  larger  for  a  given  periphe- 
ral speed  than  in  a  De  Laval  turbine.  Also  because  there  is  a 
considerable  drop  in  pressure  in  every  row  of  blades,  and  con- 
sequently a  difference  in  pressure  between  the  two  sides  of  every 
row,  there  is  always  a  leakage  of  steam  over  the  edges  of  the 
blades,  increasing  with  the  amount  of  radial  clearance  between 
the  stationary  and  moving  parts.  It  is  a  matter  of  the  greatest 
importance,  therefore,  in  designing  turbines  of  the  Parsons  type 
to  make  radial  clearances  as  small  as  possible,  consistent  with 


154 


THE  STEAM  TURBINE 


proper  allowances  for  the  expansion  due  to  unequal  heating  of 
the  parts,*  which  in  a  turbine  with  a  large  number  of  stages 
is  a  very  important  consideration.  Fig.  91  is  a  section  of  a 
typical  Parsons  rotor  and  casing  showing  by  arrows  the  leakage 
spaces  for  steam  through  the  radial  blade  clearances  a  and  b. 


FIG.  91.     Section  of  a  Typical  Parsons  Rotor  and  Casing  Showing  the  Radial 
Blade  Clearances. 


\      A  section  of  one  of  the  simplest  Parsons  turbines  is  illustrated 

1  in  Fig.  92.     The  turbine  rotor  consists  of  a  long  drum  of  three 

different  sections  supported  on  the  two  bearings  —  one  at  each 

end.     The  moving  blades  are  mounted  on  the  circumference  of 

*  Aside  from  the  question  of  radial  clearance,  all  other  points  affecting  the  design 
are  of  minor  importance  as  regards  economical  and  satisfactory  operation.  The 
most  successful  design  of  a  Parsons  type  is  the  one  which  operates  successfully 
with  the  smallest  radial  clearances.  Unequal  expansion  of  the  different  parts  of 
the  casing  and  drum  introduces  factors  which  are  very  difficult  to  estimate.  If  the 
blades  are  made  of  different  materials  from  the  drum,  at  some  temperatures  they 
are  likely  to  be  loose. 


COMMERCIAL  TYPES 


155 


156  THE   STEAM   TURBINE 

this  drum  and  the  stationary  blades  are  fitted  in  similar  rings  to 
the  inside  of  the  turbine  casing. 

The  annular  space  A  is  the  steam  chest  which  receives  high- 
pressure  steam.  The  steam  passes  through  the  alternate  rows 
of  moving  and  stationary  blades  of  the  first  section  of  the  rotor, 
through  a  second  annular  space  to  the  blades  of  the  second  sec- 
tion which  discharge  into  a  still  larger  annular  space,  from  which 
it  passes  through  the  blades  of  the  last  section  of  the  rotor  to 
the  exhaust  B.  At  the  second  and  third  annular  spaces,  where 
the  diameter  of  the  drum  is  increased,  an  unbalanced  pressure 
or  thrust  toward  the  right  is  produced  by  the  pressure  of  the 
steam;  and  this  thrust  is  increased  by  the  expansion  of  the  steam 
in  the  moving  blades  (see  Fig.  34).  To  balance  this  axial  pres- 
sure, three  balance  pistons  are  provided  at  the  left-hand  end  of 
the  casing  —  one  for  each  section  of  the  rotor.  The  smallest 
balance  piston  is  made  just  large  enough  to  equilibrate  the  thrust 
due  to  the  blades  of  the  first  section;  the  intermediate  piston 
balances  the  thrust  on  the  second  annular  area  and  that  due  to 
the  blades  of  the  second  section;  and  the  largest  piston  equili- 
brates the  pressure  on  the  third  annular  area  and  the  thrust  in 
the  third  section.  Steam  passages  are  cored  out  in  the  casing, 
as  shown  in  the  figure,  to  make  each  balance  piston  communi- 
cate with  its  corresponding  section  of  the  rotor,  so  that  the  pres- 
sure in  the  section  is  always  the  same  as  that  acting  on  the 
corresponding  balance  piston.  In  some  designs  these  cored-out 
passages  are  replaced  by  pipes  on  the  outside  of  the  casing. 
Small  annular  grooves  are  usually  cut  in  the  balance  pistons  to 
join  with  similar  annular  projections  in  the  casing.  This  con- 
struction, called  a  labyrinth  packing,  makes  the  steam  path  so 
devious  as  to  effectually  prevent  undue  leakage  of  steam  around 
the  balance  pistons. 

Oil  is  supplied  to  the  bearings  under  pressure  by  the  small 
pump  shown  in  the  figure. 

The  position  of  the  moving  blades  with  respect  to  the  stationary 
blades  (axial  clearance)  is  usually  adjusted  by  means  of  a  thrust 
or  adjustment  bearing  at  the  extreme  left-hand  end  of  the  tur- 


COMMERCIAL  TYPES 


157 


bine.  It  consists  of  a  number  of  rings  or  collars  turned  in  the 
steel  shaft  into  which  corresponding  brass  rings  in  the  adjust- 
ment bearing  are  fitted.  The  upper  and  lower  halves  of  this 
bearing  are  adjustable  and  are  moved  by  the  screws  shown  in 
the  figure.  If  the  lower  half  of  the  bearing  is  set  so  that  the 
collars  on  the  shaft  are  in  contact  on  their  left  side,  the  upper 


FIG.  93. 


Propeiter  of  a  Water-packed  Gland  of  a  Westinghouse-Parsons 
Turbine. 


half  would  have  the  collars  in  contact  on  the  right  side.  By 
this  means,  when  the  bearing  is  once  set,  the  rotor  cannot  move 
an  appreciable  distance  either  to  the  right  or  to  the  left.  A 
typical  adjustment  bearing  is  shown  more  clearly  at  the  right 
in  Fig.  108.  In  this  design  the  upper  and  lower  halves  are 
moved  by  micrometer  screws,  so  that  the  axial  position  of  the 
rotor  is  indicated  at  all  times  by  the  dials  on  these  adjusting 
screws. 

In  Fig.  92  a  very  common  method  for  operating  the  governor 


158 


THE  STEAM  TURBINE 


of  steam  turbines  is  illustrated.  A  worm  gear  on  the  main  tur- 
bine shaft  engages  with  a  gear  wheel  which  by  means  of  other 
gears  rotates  the  governor  shaft. 

Packing  Glands.  In  every  turbine,  glands  or  stuffing-boxes 
must  be  provided  where  the  shaft  passes  through  the  ends  of  the 
casing  to  prevent  the  escape  of  steam  at  the  high-pressure  end 
and  the  entrance  of  air  at  the  low-pressure  end  of  condensing 
turbines.  Steam-packed  glands  of  various  types  are  often 
provided ;  but  in  the  Westinghouse-Parsons  turbine  water-packed 
glands  are  now  generally  used.  This  arrangement  consists  of 
the  propeller  of  a  centrifugal  pump  (Fig.  93)  which  rotates  in 
the  water  supplied  to  an  annular  groove  in  the  casing.  When 


1  Stationary  Blades 


2    Moving 
•       Blades 


f 


Stationary  Blades 


Moving 
Blades 


FIG.  94.    Typical  Blading  of  a  Parsons  Turbine. 

the  turbine  is  operating  the  water  is  thrown  outward  by  the  vanes 
and  completely  fills  the  space  around  the  periphery  of  the  pro- 
peller. By  this  means  the  leakage  of  steam  or  air  is  effectually 
prevented.  As  there  are  no  rubbing  surfaces  in  these  glands  and 
no  oil  is  used,  there  is  no  contamination  of  the  exhaust  steam. 

Blades.  The  shape  and  relative  position  of  the  moving  and 
stationary  blades  in  a  Parsons  turbine  are  shown  clearly  in 
Fig.  94.  Stationary  blades  are  shown  by  cross-hatched  sections, 
and  moving  blades  by  shaded  sections. 

The  blades  of  Westinghouse  turbines  are  secured  to  the  rotor 
by  means  of  slots  turned  on  its  periphery,  which  are  slightly 


COMMERCIAL  TYPES 


159 


narrower  at  the  top  than  at  the 'bottom.  Into  these  slots  the 
blades  which  have  been  cut  at  the  roots  to  fit,  are  put  singly. 
Soft  metal  spacing  pieces  of  the  required  shape  to  fill  the  space 
in  the  slot  between  the  blades  are  calked  to  hold  the  blades 
firmly  by  a  dovetail  construction.  This  construction  is  required 
for  the  attachment  of  the  moving  blades  to  give  the  necessary  sup- 
port against  centrifugal  forces;  but  as  the  stationary  blades,  which 
are  fixed  to  the  inside  of  the  casing,  are  not  subjected  to  centrifugal 
forces,  the  slots  for  these  blades  are  not  usually  dovetailed. 


FIG.  95.     Blades  on  the  Rotor  of  a  Westinghouse  Turbine. 

Blade  Lashing  and  Shroud  Rings.  It  has  been  found  necessary 
to  bind  the  blades  together  at  their  ends  to  make  a  stronger  con- 
struction. In  the  earlier  designs  of  Parsons  turbines  the  blades  were 
usually  bound  together  with  wires  soldered  to  their  ends.  Some- 
times, however,  the  blades  were  turned  over  at  their  outer  ends  to 
form  flanges  which  were  soldered  together  into  a  solid  shroud. 

Fig.  95  shows  several  rows  of  blades  of  a  Westinghouse  turbine. 


i6o 


THE  STEAM  TURBINE 


All  blades  more  than  two  inches  long  are  reenforced  by  lashing 
with  a  wire  of  special  section  threaded  through  punched  holes 
in  the  ends  of  the  blades.  This  method  of  lashing  is  illustrated 


¥ 


Sfl/Y!t*LCS    OF  THE. 

LR5HWG 


FIG.  96.     Method  of  Lashing  Westinghouse  Blades. 

by  Fig.  96.  The  lashing  wire,  which  is  drawn  to  have  a  cross- 
section  resembling  a  comma,  binds  the  blades  together  firmly 
enough  to  give  adequate  strength  for  normal  service,  yet,  unlike 
a  very  rigid  blade  construction,  it  will  yield  in  emergencies  without 


COMMERCIAL  TYPES  l6l 

seriously  damaging  other  parts  of  the  turbine.  The  blades  are 
lashed  in  sections  three  feet  long.  Because  of  the  peculiar  shape 
of  the  section  of  the  lashing  wire,  it  can  be  calked  at  the  end  so 
that  a  "key"  remains  in  the  punched  hole  to  prevent  the  blade 


x   \  ALLIS-  CHALMERS    CO 

PATENTED  xx 
FIG.  97.     Sankey's  Bidding  for  Parsons  Turbines. 

from  getting  out  of  line.     In  many  respects  it  is  practically  as 
effective  as  a  shroud  ring. 

A  type  of  blading  for  Parsons  turbines,  patented  by  H.  R. 
Sankey  in  1903,  has  been  applied  with  certain  modifications  in 
the  Allis-Chalmers  and  the  Willans  turbines.  A  typical  illus- 
tration of  this  blading  is  shown  in  Fig.  97.  It  is  distinguished 


1 62  THE  STEAM  TURBINE 

principally  from  the  usual  Parsons  blading  by  the  attachment  of 
a  U-shaped  shroud  ring,  B,  around  both  the  moving  and  the 
stationary  blades. 

The  blades  are  cut  to  the  required  length  from  bars  of  copper 
alloy  drawn,  like  wire,  to  a  suitable  shape.  After  the  blades 
are  cut  from  the  bar,  they  are  formed  in  machine  tools  of 
special  design,  so  that  at  the  root  they  have  an  angular  "dove- 
tail" shape  as  illustrated  in  the  figure,  where  the  blades  are  shown 
inserted  in  a  suitable  foundation  ring,  A.  After  this  foundation 
ring  is  turned  to  the  proper  diameter,  " dovetail"  slots  for  the 
blades  (see  Fig.  98)  are  cut  by  a  special  milling  machine 
intended  for  very  accurate  spacing  and  inclination  so  as  to  give 
the  required  pitch  and  angle  to  the  blades. 


FIG.  98.    Spacing  for  Sankey's  Blading. 

After  the  roots  of  the  blades  have  been  inserted  in  the  founda- 
tion rings,  which,  in  cross-section,  are  also  of  a  dovetail  shape, 
the  rings  are  inserted  into  corresponding  grooves  in  the  drums 
of  the  rotor  and  in  the  inside  of  the  casing  where  they  are  held  in 
place  by  "key  pieces."  Each  of  these  "key  pieces"  after  being 
driven  into  place  is  upset  in  an  undercut  groove  which  serves  as  a 
locking  device.  The  dovetail  shapes  used  in  this  construction 
make  the  attachment  of  the  blades  at  their  roots  very  secure. 

The  channel-shaped  shroud  rings  are  purposely  made  thin  at 
the  flanges  so  that  in  case  of  contact  between  the  revolving  and 
stationary  parts  these  flanges  will  be  worn  off  at  their  edges 
without  tearing  out  or  bending  the  blades.  By  this  method,  as 
well  as  with  all  other  types  of  shroud  ring  construction,  the  strength 
of  the  blading  depends,  not  on  the  strength  of  a  single  blade, 
but  on  the  total  strength  of  as  many  blades  as  are  bound  together. 
In  the  Allis-Chalmers  turbines  all  the  blades  in  a  semi-circum- 


COMMERCIAL    TYPES  163 

ference  are  joined  by  a  shroud  ring.     The  blading  is  thus  made 
up  in  half  rings,  which  are  made  almost  entirely  by  machinery. 


FIG.  99.     Interior  of  an  Allis-Chalmers  Turbine  Casing,  Showing   Blades  Pro- 
tected by  Shroud  Rings. 

Each  ring  can  be  thoroughly  inspected  before  being  placed  in  the 
turbine  and  the  possible  inaccuracies  of  hand  work  are  likely  to 


164 


THE   STEAM  TURBINE 


be  eliminated.     Fig.  99  shows  the  interior  of   the  casing  of  a 
turbine  fitted  with  shroud  rings  on  the  blades. 

If  small  radial  clearances  are  desired,  exceptional  precautions 
in  designing  must  be  taken  to  avoid  unequal  expansions  of  the 
parts  of  the  rotor,  the  casing,  and  the  blades,  because  shroud 
rings  in  reaction  turbines  are  liable  to  produce  disastrous  results 
by  "stripping"  the  blades.  Usually  in  case  of  accident,  however, 
damaged  or  worn  rings  can  be  removed  and  the  turbine  continued 
in  operation  until  they  can  be  replaced. 


FIG.  100.     A  Westinghouse  High  Speed  Flexible  Bearing. 

Bearings.  In  turbines  of  the  Parsons  type  operating  at  above 
1800  revolutions  per  minute,  a  design  of  flexible  bearing  (Fig.  100) 
is  used  to  reduce  the  vibrations  of  the  shaft  by  permitting  the 
rotor,  when  passing  its  critical  speed,  to  revolve  about  its  center 
of  gravity  instead  of  its  geometric  axis.  This  flexible  bearing 
consists  of  a  nest  (usually  four)  of  loosely  fitting  cylindrical 
bronze  sleeves  between  which  oil  films  are  maintained  by  capil- 


COMMERCIAL  TYPES  165 

lary  attraction.*  The  clearance  between  these  sleeves  is  about 
.004  inch.  These  films  of  oil^  have  also  a  cushioning  effect  in 
absorbing  vibrations  that  occur  when  bringing  the  turbine  up  to 
speed.  This  flexible  bearing  accomplishes  the  same  purpose  for 
which  De  Laval  used  a  flexible  shaft.  In  the  figure  the  outer 
casing  of  the  bearing  is  at  the  right-hand  side  and  the  holder 
for  the  Babbitt  metal  lining  and  the  cylindrical  sleeves  around 
it  are  shown  at  the  left. 

In  larger  machines  which  run  at  lower  speeds,  balancing  is  less 
difficult  and  single  spherical-seated  bearings  lined  with  Babbitt 
metal  are  used.  Quadrant  liners  are  provided  for  either 
type  of  bearings  to  accurately  adjust  the  rotor  to  a  central 
position. 

Stages.  In  this  type  of  turbine  low  blade  speeds  are  secured 
by  using  a  larger  number  of  stages.  Thus  in  a  4oo-kilowatt 
Westinghouse-Parsons  turbine  there  are  58  stages  or  116  rows  of 
blades.  In  such  a  turbine  there  are  about  30,000  blades.  It  is 
important  to  notice  why  the  pressure  difference  for  each  row  of 
blades  gradually  decreases  from  the  admission  to  the  exhaust  in 
such  a  turbine.  Since  there  are  58  stages,  if  the  pressure  dif- 
ferences were  made  equal  for  a  total  drop  in  pressure  of  say  from 
175  pounds  per  square  inch  to  i  pound  per  square  inch,  the  drop 
in  pressure  in  each  stage  would  be  3  pounds  per  square  inch. 
But  because  the  steam  velocity  for  a  given  difference  in  pressure 
is  very  many  times  as  great  at  i  pound  as  at  175  pounds,  such  a 
division  is  not  desirable,  and  instead  the  pressure  drop  is  made  to 
give  nearly  constant  velocity  in  the  different  rows  of  blades.  A 
good  rule  for  designers  is  to  set  150  and  450  feet  per  second 
respectively  for  the  minimum  and  maximum  velocities  at  the 
high-pressure  end,  and  500  to  600  feet  per  second  at  the  low- 
pressure  end. 

A  large  Westinghouse-Parsons  turbine  is  shown  in  Fig.  101, 
with  the  upper  half  of  the  casing  removed  to  show  the  rotor, 

*  Bearing  pressure  in  pounds  per  square  inch  times  peripheral  velocity  of  the 
shaft  in  feet  per  second  is  generally  about  2500.  —  Proc.  Inst.  Elec.  Engrs.,  June, 
1905. 


i66 


THE  STEAM  TURBINE 


& 


J 


COMMERCIAL  TYPES  167 

blades,  and  balance  pistons.  The  collars  on  the  balance 
pistons  which  form  the  labyrinth  packing  are  plainly  visible. 
The  increasing  length  of  the  blades  of  the  third  (exhaust) 
section  is  also  very  apparent. 

Besides  the  Westinghouse  Machine  Company  of  Pittsburg, 
Pa.,  other  important  manufacturers  of  Parsons  turbines  are  the 
following : 

Allis-Chalmers  Company,  Milwaukee,  Wis. 

C.  A.  Parsons  &  Co.,  Newcastle,  England. 

Willans-Robinson  Company,  Rugby,  England. 

Brown-Boveri    &    Co.,    Baden,   Switzerland,   and    Mannheim, 

Germany.* 
British  Westinghouse  Company,  Manchester,  England. 

The  Allis-Chalmers  steam  turbine  is  a  reaction  type  which 
differs  from  the  original  Parsons  machines  principally  in  manu- 
facturing details  intended  to  remove  some  of  the  operating  diffi- 
culties of  the  older  designs.  An  innovation  in  the  design  of  these 
turbines  is  in  the  arrangement  and  construction  of  the  balance 
pistons.  In  the  older  types  of  reaction  turbines  the  three  balance 
pistons  were  put  at  the  high-pressure  end  of  the  turbine.  Some- 
times, however,  there  was  difficulty  with  this  construction,  as 
the  largest  or  low-pressure  piston  in  large  turbines  was  of  com- 
paratively large  diameter,  so  that  an  inner  web  was  required  in 
its  construction.  This  web  sometimes  tended  to  warp  so  as  to 
bring  the  "dummy"  or  baffle  rings  of  the  labyrinth  construction 
on  these  pistons  into  contact  with  those  attached  to  the  casing. 
To  overcome  this  difficulty  the  largest  balance  piston  has  been 
placed  at  the  low-pressure  end  of  the  rotor  behind  the  last  row 
of  blades.  In  this  location  its  effective  area  starts  from  a  smaller 
inner  diameter,  so  that  the  required  area  can  be  obtained  with  a 
smaller  outer  diameter. 

Fig.     102     represents    diagrammatically    an    Allis-Chalmers 

*  A  24,000-horsepower  steam  turbine  has  been  constructed  at  the  Mannheim 
works  of  Brown-Boveri  &  Co.  for  the  Krupp  steel  works  and  blast  furnace  plant  at 
Rheinhausen.  It  is  probably  the  largest  turbine  yet  ordered  for  stationary  service. 

The  governing  and  overload  valve  designs  of  Brown-Boveri  &  Co.'s  turbines 
are  described  and  discussed  on  pages  232  and  241. 


1 68 


THE  STEAM  TURBINE 


COMMERCIAL  TYPES 


169 


o' 


I/O 


THE  STEAM  TURBINE 


design  of  Parsons  turbine.  There  are  three  sections  of  the 
rotor  —  H,  J,  and  K  —  and  three  corresponding  balance  pistons, 
L,  M,  and  Z.  The  construction  of  the  rotor  of  one  of  these 
turbines  is  shown  in  Fig.  209.  Steam  admission  valves  are 
shown  as  in  the  usual  Parsons  designs.  The  valve  D  admits 
steam  to  the  high-pressure  end  of  the  turbine  and  is  always  under 
the  direct  control  of  the  governor.  The  second  valve,  V,  called 
the  overload  valve,  is  opened  only  when  the  turbine  must  be 
operated  at  overload  or  non-condensing  when  the  condenser 
equipment  is  out  of  service  (see  page  243).  At  C  the  main  steam 
pipe  enters  the  steam-chest  and  the  exhaust  is  at  G.  Main  bear- 
ings are  at  A  and  B. 

A  55oo-kilowatt  Allis-Chalmers  turbine-generator  is  illustrated 
in  Fig.  103. 

Governors  and  Low-Pressure  Turbines.  The  various  methods 
for  governing  Parsons  turbines  and  the  designs  of  low-pressure 
steam  turbines  of  the  Parsons  type  are  discussed  in  Chapters 
VIII  and  IX. 


r 

s 
&» 

5 

10 

*— 

—  * 

—  — 

— 

—  -• 

— 

-^ 

•i— 

__ 

1000        2000        3000        4000        5000        COOO        70 

Rated  Full  Load  Output— Kw. 

FIG.  104.     Approximate  Steam  Consumption  of  Any  Size  of  Parsons  Turbine. 

Economy  Curves.  Fig.  104  shows  fair  average  values  of  the 
steam  consumption  of  good  designs  of  Parsons  turbines  for  165 
pounds  per  square  inch  absolute  steam  pressure,  28  inches 
vacuum,  and  o°  F.  superheat.  American  Parsons  turbines,  until 
recently,  were  not  made  in  smaller  sizes  than  400  kilowatts. 
Typical  tests  and  load  curves  of  300,  500,  and  1000  kilowatt 
Westinghouse-Parsons  turbines  are  given  on  pages  268  and  269. 

The  curves  in  Fig.  105  are  based  upon  the  results  of  tests  of  a 
Westinghouse-Parsons  steam  turbine  of  standard  construction. 


COMMERCIAL  TYPES 


I/I 


It  is  stated  by  the  manufacturers  that  the  performance  as  shown 
by  these  curves  is  typical  of  machines  of  this  type. 

The  diagonal  lines  or  "Willans  lines"  in  the  figure  show  the 
total  water  weighed  or  steam  condensed  per  hour  at  various  loads. 
The  curves  or  "  water  rate  curves"  show  the  variation  in  water, 
or  more  correctly,  in  steam  consumption  per  horsepower-hour  at 


1000  1600 

LOAD    IN    KILOWATTS 


FIG.  105.     Typical   Economy  Curves  of  a  looo-Kilowatt  Westinghouse-Parsons 

Steam  Turbine. 

various  loads,  that  is,  the  "water  or  steam  rate"  of  the  turbine. 
Each  "  water  rate  curve"  corresponds  to  a  "  Willans  line"  --the 
upper  curve  to  the  upper  line,  the  lower  curve  to  the  lower  line,  etc. 
Operating  conditions  of  these  tests  are: 

(1)  Condensing  —  saturated  and  superheated  steam  (ioo°F.) 

(2)  Non-condensing  —  saturated  and  superheated  steam  (100° 

p.). 

(3)  One-quarter  rated  load  to  100  per  cent,  overload. 


172 


THE   STEAM  TURBINE 


In  the  two  overload  tests  the  operation  of  the  automatic 
secondary  or  overload  valve  may  be  observed.  As  before  noted, 
it  comes  into  action  at  a  definite  predetermined  load  as  indicated 
by  a  bend  in  the  water  line.  With  the  aid  of  this  valve  the  best 
economy  of  the  turbine  is  secured  throughout  the  range  of  normal 
loading,  while  large  overload  capacity  is  available  when  desired, 
although  at  slightly  decreased  efficiency.  When  the  secondary 
valve,  however,  has  come  fairly  into  action,  the  efficiency  under- 
goes gradual  improvement,  as  shown  by  the  reversal  of  curvature 
of  the  curves  of  steam  consumption. 


V 

\ 

V«o 

* 

Effect  of  Vacuum  and  Superheat 
on  Steam  Consumption 
1500  K.W.  Turbine    Full  Load 

x^ 

I   • 

• 

^N. 

\ 

150  Ll 
28 

>s.  Stea 
Inches 

m  Pres 
Vacuui 

sure 
l 

N 

S^ 

150Lbs.Stei 
Dry  Satur 

m  Pres 
ited  St 

-sure 
am 

V 

^ 

tj 

^ 

\^ 

^ 

16 


tt  15.5 


H 

w  15 


- 


13.5 


13 


12.5 


12 


Vacuum  -  Inches 


0        20 


FIG.  106. 


40    60    80   100   120  140 
Superheat  Deg.  F. 

Curves  of  Steam  Consumption  of  a  i5oo-Kilowatt  Westinghouse 
Turbine  with  Varying  Vacuum  and  Superheat. 


A  turbine  designed  for  condensing  work  will  not  operate  non- 
condensing  with  quite  as  good  economy  as  if  designed  to  exhaust 
against  atmospheric  pressure.  That  this  economy  is,  however, 
excellent  is  shown  by  the  upper  pair  of  curves.  The  water  rate 
is  somewhat  less  than  double  the  condensing  water  rate. 

Fig.  1 06  illustrates  graphically  the  effect  of  vacuum  and  super- 
heat on  the  steam  consumption  of  a  i5oo-kilowatt  Westinghouse 
turbine.  The  percentage  change  in  the  steam  consumption  is 
said  to  be  about  the  same  for  all  sizes. 


COMMERCIAL  TYPES  173 

THE   WESTINGHOUSE    "IMPULSE   AND    REACTION"  DOUBLE- 
FLOW  TURBINES. 

The  double-flow  principle  has  been  adopted  recently  for  the 
design  of  large  sizes  of  Westinghouse  turbines  largely  for  mechan- 
ical reasons  —  primarily  to  avoid  the  end  thrust  which  is  an  impor- 
tant factor  in  all  reaction  types.  In  small  machines,  however, 
the  double-flow  principle  does  not  have  the  same  advantages  as 
in  the  large  machines.  It  is  very  obvious  that  the  economy  of 
two  small  machines  is  not  nearly  as  good  as  one  of  twice  the 
capacity.  With  large  machines,  however,  the  change  in  econ- 
omy is  not  nearly  so  great  when  the  capacity  is  doubled.  This 
fact  is  well  illustrated  by  the  curve  in  Fig.  104.  Only  for  large 
Westinghouse  turbines  above  $000  kilowatts  capacity,  therefore, 
are  the  double-flow  designs  used;  and  for  the  smaller  sizes  the 
regular  Westinghouse-Parsons  single-flow  type  is  used  as  previ- 
ously. 

Fig.  107  illustrates  a  Westinghouse  double  flow  turbine  with 
an  impulse  element.  In  its  essential  parts  this  turbine  consists 
of  a  set  of  nozzles,  an  impulse  wheel  with  two  velocity  stages,  one 
intermediate  section,  and  two  low-pressure  sections  of  Parsons 
blading.  Steam  enters  the  turbine  through  an  opening  in  the 
lower  half  of  the  casing,  from  which*  it  is  piped  directly  to  the 
nozzle  block  shown  at  the  top  of  the  figure.  Steam  escapes  from 
these  nozzles*  at  a  high  velocity  to  impinge  on  the  impulse 
blades.  The  casing  around  the  impulse  wheel  is  made  of  suf- 
ficient size  to  permit  a  good  distribution  of  the  steam,  so  that  it 
will  enter  the  intermediate  Parsons  section  evenly  around  the 
entire  circumference  of  the  rotor.  After  the  steam  has  passed 
through  the  intermediate  section  it  divides  along  two  separate 
paths.  One  half  enters  the  left-hand  section  of  the  low-pressure 
Parsons  blading  and  the  other  half  passes  through  the  interior  of 

*  These  nozzles  are  made  non-expanding.  It  has  been  shown  that  non-expand- 
ing nozzles  give  higher  efficiencies  than  expanding  nozzles  with  steam  at  less  than 
about  70  pounds  gauge  pressure.  (See  footnote,  page  47.)  The  designers  of  these 
turbines  have  recognized  that  there  are  nozzle  losses  due  to  under-expansion  in  a 
diverging  or  expanding  nozzle  when  the  steam  is  throttled  at  light  loads. 


174 


THE   STEAM  TURBINE 


the  rotor  shell  which  forms  the  connecting  passage  to  the  right- 
hand  low-pressure  section.  Arrows  indicate  in  the  figure  the 
passage  of  the  steam  through  the  shell.  When  the  steam  is  dis- 
charged from  the  last  rows  of  low-pressure  blades,  it  passes  into 


the  exhaust  pipes  —  of  which  there  is  one  at  each  end  —  and 
then  to  the  condenser. 

As  there  is  practically  no  expansion  in  the  impulse  blades,  these 
blade  areas  are  made  to  increase  only  in  proportion  to  the  reduc- 
tion in  steam  velocity  in  each  row  of  moving  blades. 


COMMERCIAL  TYPES  175 

As  the  same  pressure  exists  on  both  sides  of  the  impulse  wheel 
disk,  this  is  not  subjected  to  any  end  thrust,  and  requires  no 
balancing.  The  small  thrust  due  to  the  difference  of  pressure 
between  the  inlet  and  outlet  of  the  Parsons  intermediate  section 
is  accurately  equilibrated  by  a  "dummy"  or  balance  piston,  of 
moderate  dimensions,  located  between  the  impulse  wheel  and 
the  right-hand  low-pressure  section.  The  thrusts  in  the  low- 
pressure  sections  are  in  opposite  directions,  and  are  therefore 
balanced.  With  these  arrangements  it  is  possible  for  the  entire 
turbine  to  run  in  perfect  equilibrium  under  all  conditions  of 
vacuum,  pressure,  and  load.  It  is,  of  course,  necessary  to  provide 
means  for  accurately  fixing  the  axial  position  of  the  rotor,  and 
for  this  purpose  an  adjustment  bearing,  shown  at  the  right-hand 
end  of  the  shaft  in  Fig.  107,  is  provided.  It  consists  of  a  number 
of  collars  turned  in  the  steel  shaft,  into  which  fit  corresponding 
brass  rings  fixed  in  the  adjustment  blocks.  The  upper  and 
lower  halves  of  the  adjustment  bearing  may  be  moved  by  means 
of  micrometer  screws,  thus  permitting  the  axial  position  of  the 
rotor  to  be  accurately  known  at  all  times. 

All  double-flow  cylinders  are  made  in  two  parts,  the  upper  and 
lower  halves  each  being  a  one-piece  casting.  The  design  is 
symmetrical  throughout,  devoid  of  longitudinal  flanges  except 
those  at  the  center  required  for  bolting  the  two  parts  together. 
The  castings  are  first  rough-bored,  after  the  flanges  have  been 
planed  and  drilled,  and  are  then  " seasoned"  with  high-pressure 
steam  for  a  number  of  hours  to  remove  any  local  casting  stresses 
in  the  metal.  They  are  then  given  the  finishing  cut  and  assembled 
with  the  boring  bar  running  in  the  bearing  housing  so  as  to 
insure  a  concentric  bore.  Manholes  are  provided  at  each  end 
of  the  cylinder  to  permit  access  for  interior  examination,  and 
auxiliary  relief  valves  are  fitted  in  each  of  the  manhole  covers  to 
prevent  the  pressure  in  the  exhaust  passages  from  rising  to  a 
dangerous  point  in  case  of  failure  of  the  condensing  apparatus 
or  the  sticking  of  the  atmospheric  relief  valve  in  the  exhaust 
piping. 

A  Y-connection,  fitted  with  two  corrugated  copper  expansion 


THE   STEAM  TURBINE 


joints  located  below  the  base  of  the  turbine,  connects  the  sepa- 
rate exhausts  to  the  main  exhaust  pipe.  These  expansion  joints 
provide  for  the  desired  freedom  of  movement  of  the  turbine  casing 
due  to  expansion  and  contraction.  An  atmospheric  exhaust 
valve  at  the  side  of  the  exhaust  "  Y"  can.be  opened  to  permit  non- 
condensing  operation. 

The  rotating  element  of  the  turbine  is  built  up  of  five  cast-steel 
parts,  in  addition  to  the  shaft.     As  may  be  seen  in  Fig.  107, 


FIG.  108.     7500- Kilowatt  Westinghouse  Turbine. 

these  are  the  three  Parsons  blading  supports,  the  impulse  section, 
and  a  dished  plate.  It  is  stated  that  the  shaft  carries  its  load  at 
one-third  the  distance  from  the  points  of  support,  so  that  this 
design  allows  a  lighter  shaft  than  would  be  required  for  dis- 
tributed loading,  and  the  consideration  of  deflection  is  practically 
eliminated.  This  built-up  part  of  the  rotor  is  rigidly  attached 


COMMERCIAL  TYPES  177 

to  the  shaft  only  at  the  right-hand  support,  and  the  opposite  end 
is  fitted  with  a  bronze  bushjng  surrounding  the  shaft,  so  as  to 
permit  the  rotor  to  move  axially,  without  appreciable  resistance, 
under  any  differential  expansion  of  shaft  and  rotor.  The  impulse 
section  consists  of  a  flanged  cast-steel  disk  forced  on  the  body 
carrying  the  intermediate  Parsons  blading.  The  flange  of  this 
disk  is  grooved  at  the  base  and  forms  the  dummy  piston  for 
balancing  the  thrust  of  the  intermediate  Parsons  section.  Fig. 
1 08  is  a  half-tone  illustration  of  a  75oo-kilowatt  Westinghouse 
double-flow  turbine. 

The  rotor  of  a  6ooo-kilowatt  double-flow  turbine  is  shown  in 
Fig.  109.     Details  of  the  arrangement  of  nozzles  and  blades  are 


FIG.  109.    Rotor  of  a  Westinghouse  Double-Flow  Turbine. 

shown  in  Fig.  no.  It  is  seen  that  the  nozzle  block  is  a  casting 
quite  separate  from  the  turbine  casing.  As  it  receives  steam 
from  the  governor  valve  this  restricts  the  high  pressure  and  high 
temperature  to  a  comparatively  small  casting  which  is  free  to 
expand  and  contract  with  changes  of  temperature. 

A  new  type  of  shaft  coupling  for  Westinghouse  turbines  is 
illustrated  in  Fig.  in. 

Westinghouse  Emergency  Speed  Limit.  A  very  interesting 
mechanism  is  provided  with  Westinghouse  turbines  for  shutting 
off  the  steam  supply  in  case  the  governor  fails  to  act  and  a 
dangerous  speed  might  be  attained.  Details  of  this  mechanism 
are  shown  in  Figs.  ii2a  and  nib.  In  its  essential  elements  it 


THE  STEAM  TURBINE 


FIG.    no.      Westinghouse  Nozzle  Block,  Showing  Arrangement  of  Nozzles  and 

Blades. 


FIG.  in.     Westinghouse  Shaft  Coupling. 


COMMERCIAL  TYPES 


179 


consists  of  a  "  weight  pin  "  P,  placed  diametrically  at  right  angles 
to  the  axis  of  the  shaft,  in  a  cylindrical  "body"  screwed  on  the 
main  turbine  shaft  at  the  high-pressure  end.  Centrifugal  force 
tends  to  drive  this  pin  away  from  the  center  and  through  the 
loosely  fitting  collar  N.  This  force  is  resisted,  however,  by 
the  "  weight  spring "  shown  around  the  pin  in  the  figures. 
The  strength  of  this  spring  can  be  adjusted  by  means  of  the 


COVER       - 
TRIGGER  CAM  BUSHING •\>ji5i£«} 


TRIGGER  CAM 
TRIGGER  CAM  BUSHING) 

TRIGGER 
WEIGHT  SPRING  RETAINER  LOCK 


BALANCING  BLOCK  PIN 
BALANCING  BLOCK 

BODY  WEIGHT 

WEIGHT  SPRING  RETAINER 
BODY  LOCKING  SCREW 

WEIGHT  SPRING  RETAINER  LOCK  SPRING 


VALVE  LEVER 
VALVE  LEVER  PIN 
VALVE  LEVER  PLATE 

.,  TRIGGER  PLATE 
TRIGGER   STOP 
TRIGGER 


'ALVE  BODY  COVER 
-  VALVE  BODY  (UPPER) 

-  VALVE  SPRING  ADJUSTING  SCREW 
—  VALVE  SPRING  RETAINER(UPPER) 

— VALVE  SPRING 

VALVE  SPRING  RETAINER  (LOWER) 

-  VALVE  BODY  (LOWER) 

.VE 


FIG.  ii2a.    Phantom  View  of  Westinghouse  Emergency  Speed  Limit. 


collar  N,  which  is  provided  with  a  screw  thread.  Such  adjust- 
ment determines  the  speed  at  which  the  centrifugal  force  over- 
comes the  spring  and  forces  the  pin  outward  to  engage  with  a 
trigger  cam  L.  This  cam  is  rigidly  attached  to  one  end  of  a  short 
shaft  S,  which  carries  at  its  other  end  a  trigger  H.  A  small  plate 
at  the  bottom  of  the  valve  lever  C  is  supported  normally  at  one 
end  on  the  trigger  H  and  at  the  other  end  on  a  screw  provided 
for  adjusting  the  spring  on  the  auxiliary  steam  valve  E. 


i8o 


THE   STEAM  TURBINE 


If  the  speed  of  the  turbine  should  become  higher  than  the  limit 
for  which  the  "weight  spring"  is  set,  the  pin  P  is  forced  out  to 
engage  with  the  cam  L,  which  in  turn  moves  the  trigger  H  away 
from  the  valve  lever  plate  which  it  supports.  In  this  way  the 
valve  E  is  opened  because  the  tension  in  the  spring  on  its  spindle 


FIG.   ii2b.     Drawings  of  Westinghouse  Emergency  Speed  Limit. 


is  released.  There  is  always  high-pressure  steam  on  the  upper 
side  of  the  valve  E,  and  when  it  is  removed  from  its  seat  this 
steam  rushes  through  a  pipe  connecting  the  lower  side  of  the 
valve  to  a  small  cylinder  at  the  side  of  the  main  steam  pipe.  A 
short  rod  attached  to  a  piston  in  this  cylinder  is  moved  by  the 
steam  pressure  to  strike  a  trigger  which  releases  and  closes  the 
emergency  valve  on  the  main  steam  pipe. 


COMMERCIAL  TYPES  l8l 

Advantages  of  the  Westinghouse  Double-flow  Type.  In  large 
capacities  the  following  advantages  are  claimed  for  the  double- 
flow  type  over  the  usual  Parsons  designs: 

(1)  Reduction  in  size  and  weight  due  to  higher  permissible 
speed. 

(2)  Almost  negligible  end  thrust. 

(3 )  Blades  and  casing  are  not  exposed  to  steam  at  high  temper- 
atures. 

(4)  Large  volume  per  pound  of  steam  at  the  admission  to  the 
first  Parsons  section  avoids  the  use  of  very  short  blades. 

(5 )  Only  one  balance  piston  is  required  and  this  is  of  relatively 
small  diameter.  1 

(6)  Exhaust  connections  are  considerably  reduced  in  size,  due 
to  divided  flow. 

(7)  The  impulse  element  is  well  suited  to  high  pressure  and 
superheat,  and  by  this  modification  the  shaft  length  is 
reduced  nearly  50  per  cent. 

An  exact  reproduction  of  a  section  drawing  of  a  Westinghouse 
double-flow  low-pressure  turbine  rated  at  1000  kilowatts  is 
shown  in  Fig.  184  in  Chapter  IX. 

The  following  figures  1120,  d,  e,  and  f  show  designs  used  by 
the  Westinghouse  Machine  Company  for  25,ooo-kilowatt  turbine- 
generators  to  operate  at  200  pounds  per  square  inch  absolute 
steam  pressure,  29  inches  vacuum  and  200°  F.  of  superheat. 
Fig.  ii2c  shows  the  double-flow  turbine  designed  for  these  con- 
ditions and  operating  at  1500  r.p.m.  Fig.  ii2d  shows  a  similar 
design  operating  at  750  r.p.m.  A  tandem  compound  arrange- 
ment operating  at  750  r.p.m.  is  shown  in  Fig.  ii2e.  In  this  last 
design  it  will  be  observed  that  the  high-pressure  portion  is  of 
the  ordinary  single-flow  arrangement  while  the  low-pressure 
end  is  made  double-flow.  The  combined  unit  is  connected  to 
a  single  25,ooo-kilowatt  generator.  A  cross-compound  turbine 
arrangement,  Fig.  ii2f,  with  the  high-pressure  portion  operating 
at  1500  r.p.m.  and  connected  to  a  i2,5oo-kilowatt  generator  is 
shown  here  together  with  a  low-pressure  portion  which  is  of  the 
double-flow  arrangement  and  operating  at  750  r.p.m.  The  low- 


i8ia 


THE    STEAM    TURBINE 


pressure  portion  is  also  connected  to  a  i2,5oo-kilowatt  generator. 
These  figures  show  the  comparison  to  scale  of  the  arrangements 
described  above. 

All  of  these  designs  would  give  very  excellent  economy,  and  the 
choice  of  the  unit  would  depend  primarily  on  the  two  factors  of 
first  cost  and  economy,  assuming  that  in  each  case  the  relia- 
bility for  continuous  operations  is  the  same. 


FIG.  ii2c.     Double-flow  Reaction  Turbine  Designed  for  1500  r.p.m. 

A  close  study  of  the  four  arrangements  indicates  that  the 
double-flow  turbine  at  1500  r.p.m.,  direct  connected  to  a  single 
generator  (Fig.  H2g),  is  the  cheapest  construction.  The  large 


FIG.  ii2d.    Double-flow  Reaction  Turbine  Designed  for  750  r.p.m. 

areas  required  in  the  low-pressure  stages  of  this  turbine  make 
high  velocity  and  long  length  of  blades  essential,  with  the 
necessity  of  careful  designing  to  properly  take  care  of  the  stresses 
due  to  centrifugal  force  in  the  low-pressure  end. 


COMMERCIAL  TYPES 


l8lb 


The  most  economical  combination  of  the  four  is  the  cross- 
compound  reaction  turbine  with  the  high-pressure  portion  run- 
ning at  1500  revolutions  and  the  low-pressure  portion  at  750 


FIG.  1 1 2e.     "  Tandem-compound  "  Reaction  Steam  Turbine. 

revolutions.  With  this  arrangement  the  highest  efficiency  is 
obtained,  because  the  method  of  combining  the  unit  into  highl- 
and low-pressure  cylinders,  running  at  1500  and  750  r.p.m., 


FIG.  ii2f.     "  Cross-compound  "  Reaction  Steam  Turbine. 

gives  the  condition  for  best  blading  proportions  throughout  the 
turbine  without  departing  from  established  standards  of  prac- 


i8ic 


THE    STEAM   TURBINE 


tice.  The  construction,  however,  is  considerably  heavier  than 
the  single  unit  of  the  double-flow  type  (Fig.  ii2g)  and  is  more 
costly  to  construct  and  install.  At  powers  and  speeds  attain- 
able with  single  alternating-current  units  of,  say,  below  15,000 
kilowatts'  capacity,  the  double-flow  turbine  will  be  nearly,  if  not 
equal,  to  the  cross-compound  " straight  reaction"  turbine  under 
the  same  operating  conditions,  and  any  difference  in  efficiency 
would  probably  be  offset  by  the  lower  first  cost  of  the  double-flow 
machine.  Taking  into  consideration  both  first  cost  and  effi- 
ciency, between  10,000  and  40,000  kilowatts'  capacity,  the  double- 
flow  machine  is  undoubtedly  the  proper  type  of  construction. 


FIG.  ii2g.     Section  of  Combined  Impulse  and  Double-flow  Reaction  Turbine. 

The  demand  for  turbine-generators  is  greatest  between  4000 
to  15,000  kilowatts  maximum  rated  capacity,  within  which  range 
the  double-flow  machines  of  the  combined  impulse  and  reaction 
types  of  blading  satisfactorily  meet  commercial  conditions  both 
with  respect  to  cost  and  efficiency.  There  is,  however,  practi- 
cally the  same  to  be  said  regarding  turbines  of  the  Curtis  type 
having  four  to  five  pressure  stages. 

Below  4000  kilowatts'  capacity  the  turbines  consisting  of  an 
impulse  wheel  followed  by  single-flow  reaction  blading,  and  a 
"  straight "  single-flow  reaction  turbine  (Fig.  102,  page  168, 
and  Fig.  1121),  represent  the  machines  best  suited  for  average 
operating  conditions.  In  most  cases  the  former  would  be  pre- 


COMMERCIAL   TYPES  l8ld 

f erred;  but  when  the  speed  is  to  be  made  particularly  low  the 
preference  goes  to  the  latter. 

Thus  with  a  speed  of  3600  r.p.m.  driving  a  6o-cycle  generator 
at,  say,  500  kilowatts,  the  best  design  would  be  a  combined 
impulse  and  reaction  machine  for  best  efficiency  and  lowest 
cost.  If,  however,  the  generator  to  be  driven  was  for  25-cycle 
service,  with  an  allowable  maximum  speed  of  1500  r.p.m.  and 
the  same  capacity  the  single-flow  reaction  turbine  would  be 
selected,  providing  reaction  blading  was  to  be  used  at  all. 

In  general,  however,  the  application  of  the  combined  impulse 
and  reaction  turbine,  consisting  of  an  impulse  element  for  the 
high-pressure  portion  and  reaction  blading  for  the  low-pressure 
portion,  is  well  adapted  for  complete  expansion  turbines  over 
wide  ranges  of  power  and  speed;  and  since  the  introduction  of 


FIG.  ii2h.     Relative  Lengths  of  Rotors  in  Two  Common  Types. 

this  type,  a  large  proportion  of  the  firms  building  steam  turbines 
have  utilized  this  construction,  either  with  Parsons  or  Rateau 
blading  following  the  Curtis  impulse  element  in  the  high- 
pressure  end.  The  principal  advantage  of  this  type  of  con- 
struction is  the  shortening  of  the  machine  without  very  much 
loss  in  efficiency,  the  elimination  of  balancing  pistons  with  the 
avoidance  of  the  very  considerable  leakage  of  steam  through 
them,  and  the  securing  of  high  economies  at  light  loads  by  the 
application  of  the  method  of  governing  by  "cutting  out  nozzles" 
(see  page  221),  now  applied  to  Westinghouse  turbines  of  the 
combined  impulse  and  reaction  type. 


i8ie 


THE   STEAM   TURBINE 


Fig.  H2h  shows  the  relative  lengths  of  the  rotors  of  the  latest 
design  of  Westinghouse  turbine  with  combined  impulse  blading 
and  single-flow  reaction  blading  compared  with  the  conventional 
single-flow  Parsons  type  with  "  balance  pistons  "  (see  page  156). 
Nearly  50  per  cent,  in  length  is  saved,  making  the  difficulties 
due  to  springing  and  expansion  of  the  casing  and  rotor  relatively 
small. 

Another  important  consideration  in  choosing  between  the 
"  straight "  single-flow  reaction  and  the  combined  impulse  and 
reaction  types  is  that  the  former  is  generally  preferred  for 


FIG.  ii2i.     Single-flow  Reaction  Turbine  with  Kingsbury's  Thrust  Bearing. 

moderate  superheats  and  pressures,  while  the  latter  is  selected 
when  the  superheats  and  pressures  are  high. 

Another  important  improvement  in  the  construction  of  reac- 
tion turbines  is  shown  in  Fig.  1121  where  the  Kingsbury  type  of 
thrust  bearing  is  applied.  By  this  means  the  usual  "  balance 
pistons  "  required  for  the  single-flow  type  are  eliminated. 

A  recent  design  of  a  2O,ooo-kilowatt  "  tandem  "  type  of  reac- 
tion turbine  is  shown  in  Fig.  H2J.  At  the  right-hand  side  is  the 


COMMERCIAL  TYPES 


l8lf 


FIG.  H2J.     Tandem  Reaction  Turbine. 

high-pressure  turbine  and  at  the  left  the  low-pressure.  This 
figure  shows  more  clearly  the  type  shown  also  in  Fig.  ii2e. 
There  are  very  few  applications  as  yet  for  "  land  "  service  of 
this  arrangement,  although  it  is  common  for  marine  service. 


FIG.  1 1 2k.    Relative  Sizes  of  Steam  Turbine-generators,  from  One  Kilowatt  to 
35,000  Kilowatts. 


Fig.  1 1 2k  illustrates  the  relative  sizes  of  steam  turbine-gener- 
ators for  capacities  from  one  kilowatt  for  the  smallest  to  35,000 
kilowatts  for  the  largest.  The  turbine  is  shown  on  the  right- 
hand  side  and  the  generator  on  the  left. 


1 82  THE  STEAM  TURBINE 


THE    CURTIS   TURBINE 

The  Curtis  steam  turbine,  of  which  the  original  patents  were 
issued  to  C.  G.  Curtis  about  1895,  is  manufactured  by  the 
General  Electric  Company  at  Schenectady,  N.Y.,  and  Lynn, 
Mass.,  the  British  Thomson-Houston  Company  at  Rugby, 
England,  and  the  Allgemeine  Elektrizitats  Gesellschaft  at  Berlin, 
Germany. 

As  in  the  De  Laval  turbine,  the  steam  is  expanded  in  nozzles 
before  reaching  the  moving  blades,- but  the  complete  expansion 
from  the  boiler  to  the  exhaust  pressure  occurs  in  this  type  usually 
in  a  series  of  stages  or  steps,  as  the  steam  passes  through  a  suc- 
cession of  chambers,  separated  from  each  other  by  diaphragms. 
The  diaphragms  and  blade  wheels  of  a  four-stage  Curtis  turbine 
are  shown  by  a  section  drawing  in  Fig.  113.  Each  chamber  or 
stage  contains  usually  one  disk  or  blade  *  wheel.  Steam  at  the 
admission  pressure  enters  the  first  set  of  nozzles  through  the  port 
A,  where  it  expands  to  the  pressure  in  the  first  stage  and  delivers 
a  portion  of  its  energy  to  the  blades  in  the  wheel  F.  The  steam 
then  expands  again  through  a  second  set  of  nozzles  in  the  dia- 
phragm C  leading  to  a  still  lower  pressure  in  the  second  stage, 
where  it  gives  up  a  portion  of  the  energy  remaining  to  a  second 
set  of  blades,  and  so  on.  In  the  very  small  units  but  one  pressure 
stage  is  usually  employed,  but  in  the  larger  sizes  from  two  to  five 
are  used.  The  general  arrangement  of  the  nozzles  and  blades 
in  a  single-stage  Curtis  turbine  was  shown  diagrammatically  in 
Fig.  39.  It  is  typical  of  these  turbines  that  there  are  always 
three  or  more  rows  of  blades  following  each  set  of  nozzles,  and 
at  least  one  row  is  stationary.  These  stationary  blades  are 
technically  called  intermediates.  There  is  practically  no  expan- 
sion in  the  stationary  blades;  the  object  of  the  several  rows  of 
blades  is  only  to  reduce  the  velocity,  and  for  a  given  blade  speed 

*  The  terms  vane,  blade,  and  bucket  are  often  used  interchangeably. 
Common  practice,  however,  seems  to  apply  blade  to  the  Parsons  turbine,  and 
bucket  to  the  Curtis,  De  Laval,  and  those  of  the  Pelton  type.  In  order,  however, 
that  the  notation  may  not  be  confused,  the  term  blade  will  be  used  in  connection 
with  Curtis  as  well  as  other  types. 


COMMERCIAL  TYPES 


I85 


rim.  A  dovetailing  method  similar  to  Fig.  63  is  now  generally 
preferred  to  the  method  of  inserting  the  blades  by  casting.  The 
fixed  blades,  or  intermediates,  are  also  either  cut  or  cast  in  seg- 


FIG.  115.     Curtis  Moving  Blade  Segments. 

ments  (Fig.  116),  and  are  fastened  by  bolts  to  the  interior  of  the 
casing  as  shown  in  Fig.  57.  These  intermediates  cover  only  the 
portion  of  the  circumference  upon  which  the  belt  of  steam 
delivered  by  the  nozzles  can  impinge.  To  make  the  blades  more 
rigid,  thin  bands  or  shroud  rings  are  riveted  in  segments  to  pro- 
jections on  their  ends. 


FIG.  116.     Curtis  Intermediate  Blade  Segments. 

The  wheels  of  a  four-stage  Curtis  turbine  are  shown  in  Fig. 
117.  There  are  two  rows  of  blades  on  each  wheel,  so  that  in  this 
design  there  are  two  velocity  stages  in  each  pressure  stage.  The 
shroud  rings  on  each  row  of  blades  are  plainly  visible. 

Shafts  and  Bearings.  The  smaller  sizes  of  Curtis  turbines 
have  horizontal  shafts  with  standard  bearings,  as  devices  for 
flexibility  are  unnecessary  at  the  speeds  employed.  The  larger 


1 86 


THE  STEAM  TURBINE 


COMMERCIAL  TYPES 


I87 


sizes,  however,  are  built  with  a  vertical  shaft  supported  on  an 
ingenious  step  bearing,  shown  at  the  bottom  of  Fig.  118,  which  is 
supplied  with  oil  or  water  under  pressure,  the  shaft  thus  revolving 
on  a  film  of  liquid.  The  small  disk  D  is  attached  by  dowels  E 
to  the  main  shaft.  The  bearing  is  between  the  stationary  plate 
C  and  the  disk  D.  This  vertical  shaft  arrangement  is  one  of 
the  special  characteristics  of  the  large  sizes  of  Curtis  turbines,  and 


O/J  Dra/n 
O/V  dupp/y 

FIG.  118.     Step  Bearing  for  a  Vertical  Curtis  Turbine. 

produces  a  very  compact  design.  The  direct-connected  electric 
generator  is  mounted  immediately  above  the  turbine,  as  shown  in 
Fig.  119,  which  is  a  section  of  a  pooo-kilowatt  Curtis  turbine  - 
generator. 

Fig.  120  is  a  "phantom"  view  of  a  3oo-kilowatt  Curtis  turbine- 
generator,  showing  the  wheels,  armature,  and  couplings  as  if  the 
turbine  casing  and  generator  frame  were  transparent. 

Curtis  units  are  manufactured  from  15  kilowatts   (about  20 


188 


THE  STEAM  TURBINE 


FIG.  119.     Section  of  pooo-Kilowatt  Vertical  Curtis  Turbine-Generator. 


COMMERCIAL  TYPES 


189 


FIG.   120.     Phantom  View  of  a  Curtis  Turbine  Showing  Wheels,  Armature,  and 

Couplings. 


FIG.  121.     Ring  Type  of  Emergency  Stop. 


190 


THE   STEAM  TURBINE 


horsepower)  at  3600  to  4000  revolutions  per  minute  to  as  high 
as  9000  kilowatts  (nearly  12,000  horsepower)  at  about  750 
revolutions  per  minute,  the  general  application  being  to  direct- 
connected  electric  generators  for  power  or  lighting  purposes. 

Emergency  Valve.     Since  a  steam  turbine  can  accelerate  at 
a  rapid  rate  and  this  increase  in  speed  is  not  easily  perceptible, 


am. 


FIG.  122.     Emergency  Stop  Valve. 

it  is  important  that  all  these  machines  be  equipped  with  simple 
speed  limiting  devices  which  are  operated  automatically  in  emer- 
gencies. The  device  shown  in  Fig.  121  consists  of  a  steel  ring 
(13)  placed  around  the  shaft  between  the  turbine  and  the  gener- 
ator. This  ring,  which  is  held  in  place  by  stud  bolts  (4),  is 
placed  in  a  slightly  eccentric  position,  and  the  centrifugal  force 
due  to  this  unbalancing  is  counteracted  by  a  helical  spring  (u). 


COMMERCIAL  TYPES 


191 


When  the  speed  increases,  the  centrifugal  effort  overcomes  the 
spring  and  the  ring  moves  into  a  still  more  eccentric  position  as 
indicated  by  the  dotted  lines.  In  this  position  the  ring  strikes  a 
bell-crank  lever,  which  trips,  by  means  of  a  simple  auxiliary 
mechanism  and  the  tension  rod  L  (Fig.  122),  the  throttle  valve 
on  the  main  steam  supply  pipe.  The  rod  L  is  connected  to  the 
crank  D,  which  operates  to  release  the  spring  S,  pulling  up  the 
gear  and  throwing  out  the  hook  G,  which  holds  the  valve  open. 
When  released  by  this  emergency  ring  mechanism,  the  valve 
descends  upon  its  seat  with  a  very  positive  motion  due  to  its 


FIG.  123.     Details  of  "Spring  Type"  of  Emergency  Stop. 

own  weight  and  the  unbalanced  pressure  on  the  area  of  the 
valve  stem. 

Fig.  123  shows  a  little  different  arrangement  for  tripping  the 
valve.  The  free  end  of  a  spiral  spring  is  thrown  out  by  centrifu- 
gal force  and  strikes  a  bell-crank  lever  in  very  much  the  same 
way  as  the  ring  does.  The  emergency  valve  is  opened  by  means 
of  the  hand  wheel  shown  at  the  bottom  of  the  figure. 

No  turbine  should  be  kept  in  operation  unless  it  is  known  that 
this  speed  limiting  device  is  in  reliable  condition. 

Governor.  Curtis  turbines  are  governed  by  a  method  com- 
monly known  as  "cutting  out  nozzles."  By  this  method  the 


THE   STEAM  TURBINE 


number  of  nozzles  which  are  open  for  the  discharge  of  steam  is 
regulated  according  to  the  requirements  of  the  load.  This  method 
is  described  and  typical  Curtis  governors  and  valve  gears  are 
illustrated  in  Chapter  VIII. 


FIG.  124.     25-Kilowat*  Curtis  Turbine-Generator. 

Small  Turbines.  Fig.  124  shows  a  25-kilowatt  Curtis  turbine 
and  generator  suitable  for  lighting  a  factory.  The  whole  set 
occupies  very  little  space  compared  with  that  required  for  a 
reciprocating  engine.  The  shaft,  armature,  and  turbine  wheel  of 
this  set  are  shown  separately  in.  Fig.  125.  One  of  the  latest  and 


FIG.  125.     Wheel,  Shaft,  Armature,  and  Commutator  of  a  Small  Curtis  Turbine. 

most  efficient  designs  of  blade  or  bucket  wheels  for  Curtis  tur- 
bines with  two  pressure  stages  is  shown  in  Fig.  i25a.  The 
disks  or  wheels  in  the  two  stages  are  of  the  same  diameter  but 
the  much  greater  blade  length  toward  the  low-pressure  end 


COMMERCIAL  TYPES 


193 


makes  the  actual  over-all  diameter  at  that  end  considerably 
larger  than  at  the  high-pressure  end. 


FlG.  I25a.     Latest  Construction  of  Blade  Wheels  of  Curtis  Two-stage  Turbines. 

The  most  recent  improvement  in  valve  gears  on  Curtis  tur- 
bines is  shown  on  the  turbine  illustrated  in  Fig.   125!).     The 


FIG.  1250.    Horizontal  Curtis  Steam  Turbine  with  Latest  Steam-operated  Valve 

Gear. 


194 


THE   STEAM  TURBINE 


centrifugal  governor  is  placed  at  the  upper  end  of  a  vertical  shaft 
between  the  turbine  and  generator,  which  is  driven  by  worm 
gearing  from  the  main  shaft  of  the  turbine.  The  motion  of  the 
main  governor  is  transmitted  to  the  valve  gear  by  means  of  levers 
and  rods,  which  operate  a  small  pilot  valve,  controlling  the  ad- 
mission of  steam  to  a  steam  cylinder  at  the  upper  end  of  the 
valve  mechanism.  That  is,  the  pilot  valve  serves  to  admit 
steam  either  above  or  below  the  piston  in  the  steam  cylinder. 
The  piston  rod  extends  into  the  steam  chest  and  on  this  rod 
are  mounted  a  series  of  spiders,  which  engage  a  corresponding 
series  of  annular  double-seated  admission  valves.  The  spiders 
on  the  valve  rod  are  arranged  so  that  the  valves  are  lifted  from 
their  seats  in  sequence  as  the  rod  is  raised  by  the  steam  cylinder 
under  control  of  the  pilot  valve.  As  each  of  the  valves  is  lifted 
from  its  seat,  steam  is  admitted  from  the  central  space  within 
the  annular  valves  to  cored  passages  leading  to  the  turbine 
nozzles.  Each  one  of  these  passages  in  general  communicates 
with  two  sections  of  the  turbine  nozzle. 

Correction  Curves.  Typical  curves  showing  the  variation  in 
steam  consumption  of  a  5oo-kilowatt  Curtis  turbine,  due  to  in- 
creasing superheat  and  vacuum,  are  shown  in  Figs.  126  and  127. 


r;-17 
i* 

o  £ 

II" 


20    10     60     80   100  120  140  160  180  200  220 
Superheat  Degs.  Eahr. 


22 

L 

1"  18 
17 
2 

N 

\ 

S 

\ 

\ 

\ 

2  v  •     "24              26               28               3( 

Vacuum  in  Exhaust  Base 


FIG.  126.  FIG.  127. 

Curves  Showing  the  Effect  of  Superheat  and  Vacuum  on  the  Steam  Consumption 
of  a  5oo-Kilowatt  Curtis  Turbine. 

Such  curves  become  most  useful,  however,  when  they  are  re- 
duced to  equivalent  percentages  like  those  for  De  Laval  turbines 


COMMERCIAL  TYPES 


shown  on  pages  150  and  151.     In  Chapter  VI  the  correct  method 
for  making  this  transposition  was  explained. 

Steam  Consumption.     Fig.  128  is  a  curve  to  show  approxi- 


ao 


2000 


3000  4000 

RatedJFull  Load  Output  * 


5000 


6000 


7000 


FIG.  128.  Approximate  Steam  Consumption  of  Any  Size  of  Curtis  Turbine 
with  165  Pounds  per  Square  Inch  Absolute  Pressure,  28  Inches  Vacuum,  and 
no  Superheat. 

mately  the  steam  consumption  of  any  size  of  Curtis  turbine  at  the 
rated  full  load.  All  the  data  for  this  curve  were  corrected  by 
using  percentage  curves  like  those  referred  to  above,  which  served 
to  reduce  the  conditions  of  the  various  tests  to  assumed  conditions 
of  165  pounds  per  square  inch  absolute  steam  pressure,  28  inches 
vacuum,  and  no  superheat.  To  get  sufficient  data  for  this 
curve  it  was  necessary  to  include  some  tests  made  with  com- 
mercial loads,  making  its  values  probably  a  little  higher  than 
they  would  be  if  all  the  tests  had  been  run  with  a  constant  load. 
Analysis  of  Losses  in  a  Curtis  Turbine.  Steinmetz  has  calcu- 
lated the  energy  distribution  in  a  typical  two-stage  Curtis  tur- 
bine and  has  given  the  results  in  the  form  of  the  diagram  in 
Fig.  129. 

WESTINGHOUSE   IMPULSE  TURBINES. 

Still  another  type  of  steam  turbine  intended  particularly  for 
small  capacities  has  been  developed  by  the  Westinghouse  Ma- 
chine Company,  a$  illustrated  in  Figs.  130  and  131.*  Machines 
of  this  type  are  suitable  for  a  capacity  as  low  as  one  kilowatt. 
By  this  construction  it  is  possible  to  secure  with  the  use  of  only 

*  Turbines  of  this  type  are  known  abroad  as  "  Electra  "  designs. 


THE  STEAM  TURBINE 


FIG.  129.     Analysis  of  the  Losses  in  a  Turbine  with  Three  Velocity  Stages  in 
Each  of  Two  Pressure  Stages. 


FIG.  130.     Westinghouse  Impulse  Turbine  (with  one  reversal). 


COMMERCIAL  TYPES 


1940 


one  row  of  moving  blades  an  effect  similar  to  the  velocity  stages 
in  a  Curtis  turbine  with  only  one  row  of  moving  blades  or 
buckets  as  illustrated  in  Fig/  130.  This  design  is  suitable  for 
the  pressure  drop  in  non-condensing  operation.  The  arrange- 


FiG.  131.    Westinghouse  Impulse  Turbine  (with  two  reversals). 

ment  shown  in  Fig.  131  has  two  reversals  of  the  steam  and  is 
suitable  for  condensing  operation. 


FIG.  13  la.    Three  Sizes  of  Small  Westinghouse  Impulse  Turbines. 

The  advantages  of  this  construction  are  that  it  is  essentially 
simpler  than  the  De  Laval  in  the  elimination  of  speed-reduction 
gears,  and  requires  a  very  much  smaller  number  of  blades 
than  the  Curtis  type. 


THE   STEAM  TURBINE 


COMMERCIAL  TYPES 


195 


RATEAU   TURBINES. 


Professor  Rateau  of  Paris  is' also  a  pioneer  in  the  development 
of  a  well-known  type  of  steam  turbine.  His  first  experiments 
were  made  with  a  turbine  having  a  single  impulse  wheel;  but  he 
soon  abandoned  this  type  in  favor  of  a  multiple  wheel  construe- 


.-'-:•..  y, 


FIG.  132.     Diagrammatic  Representation  of  Four  Stages  of  a  Rateau  Turbine. 

tion.  The  Rateau  turbine  is  often  called  "  multicellular," 
meaning  that  it  consists  of  a  large  number  of  "  cells  "  or 
pressure  stages  of  which  the  separating  walls  are  diaphragms 
similar  to  those  in  a  Curtis  turbine.  The  principle  of  the 
Rateau  turbine  is  illustrated  by  the  section  drawing  in  Fig.  132, 
which  shows  diagrammatically  four  stages.  Essentially  the 


196  THE   STEAM  TURBINE 

Rateau  type  differs  from  that  of  Curtis  in  that  it  has  a  much 
larger  number  of  pressure  stages  or  "cells"  but  no  velocity 
stages.  There  is  therefore  only  one  row  of  blades  in  each  stage. 
Except  for  the  fact  that  turbines  with  simple  disk  wheels  can  be 
operated  at  higher  blade  speeds  than  reaction  turbines  of  the 
drum  type  (Parsons),  making  the  efficient  utilization  of  steam  at 
higher  velocities  possible,  the  Rateau  and  the  Parsons  types 
would  require  the  same  number  of  stages.  Rateau  turbines 
have  from  20  to  40  stages  respectively,  depending  on  whether 
they  are  for  non-condensing  or  for  condensing  service.  For 
given  blade-speed,  steam  pressure,  and  superheat,  the  number  of 
stages  increases,  although  not  proportionally,  as  the  exhaust 
pressure  is  reduced. 

Nozzles  and  Diaphragms.  Annular  nozzles  are  set  in  each  of 
the  diaphragms  between  the  stages.  Because  of  the  large  num- 
ber of  stages,  the  pressure  drops  are  very  small,  so  that  the 
nozzles -are  made  with  a  uniform  cross-section  along  their  length; 
that  is,  they  are  non-expanding.  To  allow  for  the  increased 
volume  of  the  steam  as  it  expands,  in  almost  all  the  other  types 
of  impulse  .turbines  the  nozzles  are  made  with  at  least  somewhat 
larger  radial  width  for  the  lower  pressures.  In  Rateau  turbines, 
however,  the  same  increased  nozzle  area  is  secured  by  increas- 
ing only  the  arc  or  part  of  the  circumference  occupied  by  the 
nozzles.  In  the  last  stages,  then,  where  the  entire  circumference 
of  the  diaphragm  is  made  use  of,  a  complete  annular  jet 
results. 

Rateau  nozzles  are  arranged  in  groups  very  much  like  the 
Curtis  nozzle  plate  shown  in  Fig.  114.  Diaphragms  of  several 
sizes  of  these  turbines  are  shown  in  Fig.  133.  Several  groups  of 
nozzles  can  be  seen  in  each  diaphragm.  At  the  high-pressure 
end  of  .the  turbine  there  are  only  a  few  groups  (usually  about 
three),  but  in  each  succeeding  stage  there  is  a  greater  number. 
Because  the  steam  discharged  from  the  blades  is  carried  along  a 
short  distance  by  the  rotation  of  the  wheel,  a  portion  of  each 
group  of  nozzles  is  located  a  little  in  advance  of  the  preceding 
set. 


COMMERCIAL  TYPES 


197 


One  of  the  advantages  claimed  by  Rateau  for  multi-stage 
types  over  those  in  which  the  steam  is  admitted  around  the  whole 
periphery  in  all  the  stages,  is' that  since  the  volume  of  the  steam 


FIG.  133.     Diaphragms  of  a  Rateau  Turbine  Showing  Nozzles  and  a 
Shaft  Packing  with  "Water  Grooves." 

at  the  admission  end  is  small,  the  blades,  in  the  Parsons  type  for 
example,  have  necessarily  a  small  radial  height,  so  that  there  is 
more  friction  due  to  the  passage  of  steam  than  where  the  steam 
spaces  are  larger  and  the  volume  of  the  steam  is  large  in  pro- 
portion to  the  surface  of  the  blades. 


198 


THE  STEAM  TURBINE 


COMMERCIAL  TYPES  199 

Description.  Fig.  134  is  a  section  of  a  5oo-horsepower  Rateau 
turbine  of  24  stages.  It  will  'be  observed  that  this  turbine  is 
divided  into  three  sections,  the  high-pressure  and  intermediate 
sections  being  separated  from  the  low-pressure  section  by  a 
" middle"  turbine  bearing  H.  This  turbine  is  designed  to 


FIG.  135.     A  Rateau  Disk. 

operate  at  2400  revolutions  per  minute.  Dimensions  can  be 
calculated  from  the  scale  given  at  the  right-hand  side.  In  this 
figure  B  is  the  main  steam  pipe,  C  is  the  throttle  valve  controlled 
by  the  governor,  D  is  a  blade  of  one  of  the  moving  disks,  E  is  a 
nozzle  in  a  diaphragm,  F  is  a  steam  pipe  connecting  the  inter- 
mediate and  low-pressure  sections,  L  is  an  auxiliary  valve  to 


200  THE   STEAM   TURBINE 

admit  high -pressure  steam  to  "  carry  an  overload,"  and  G  is  the 
exhaust  pipe  leading  to  the  condenser.  Bushings  of  anti-friction 
metal  are  fitted  in  the  diaphragms  where  the  shaft  passes  through 
them. 

Wheel  Disks.  A  typical  Rateau  disk  is  shown  in  Fig.,  135, 
Details  of  construction  are  shown  better,  however,  in  Fig.  132. 
The  disks  are  dished  to  add  to  their  lateral  strength,  or  to 
make  them  stiffen*  In  the  recent  designs  a  shroud  ring  is 


FIG,  136.     Rateau  Disks  Assembled  on  the  Turbine  Shaft. 

fitted  around  the  blades  as  illustrated  in  the  figures.  The  blades 
resemble  those  used  in  De  Laval  and  Curtis  turbines  except  that 
they  have  a  flat  projection  at  the  root  which  is  provided  to  fasten 
them  to  the  flange  of  the  disk  by  riveting.  The  holes  shown 
in  the  disk  in  Fig.  135  were  drilled  for  balancing.  Fig.  136 
shows  a  group  of  Rateau  disks  assembled  on  the  turbine 
shaft. 

*  Diaphragms  of  large  sizes  of  Curtis  turbines  are  dished  in  the  same  way  to  give 
increased  strength.     See  Figs.  57  and  119. 


COMMERCIAL  TYPES 


201 


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i  /i 


r       PL, 

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S     o 


1 1 


202  THE    STEAM   TURBINE 

Manufacturers.  Rateau  turbines  are  constructed  by  the 
pioneers,  Sautter,  Harle  &  Co.,  at  Paris,  by  the  Maschinenfabrik 
Oerlikon  in  Switzerland  and  by  many  other  companies  in  Europe. 
American  rights  are  controlled  by  the  Southwark  Foundry  & 
Machine  Company  of  Philadelphia.  Rateau  designs  are  fre- 
quently used  in  combination  with  other  types,  as  for  example 
when  Curtis  blading  is  used  for  the  first  stage  and  Rateau  blad- 
ing  for  the  remaining  stages  (see  example,  pages  86-95) .  Licenses 
to  manufacture  Rateau  turbines  have  been  issued  to  Goldie  & 
McColloch  Company,  Gait,  Ontario,  Canada. 

A  typical  Rateau  turbine  connected  to  a  high-lift  centrifugal 
pump  is  illustrated  in  Fig.  137. 

Low-Pressure  Rateau  Turbines  are  extensively  used  in  Europe 
to  operate  with  the  exhaust  steam  from  rolling  mill  and  mine 
engines.  Professor  Rateau  has  designed  a  steam  accumulator 
(Fig.  184)  for  application  in  such  cases  where  the  steam  supply 
is  intermittent.  It  is  described  in  Chapter  IX  in  the  discussion 
of  low-pressure  steam  turbines. 

WILKINSON    TURBINES. 

Rateau  turbines  are  governed  by  throttling  the  steam  pressure 
by  means  of  valves  controlled  by  the  governor.  Mr.  James 
Wilkinson  has  invented  a  system  of  governing  steam  turbines 
(see  page  238)  which  is  intended  to  be  equivalent  to  the  Corliss 
"cut-off"  governing  of  reciprocating  engines.  He  has  applied 
this  method  of  governing,  together  with  some  other  unique 
features,  to  steam  turbines  of  the  Rateau  type  which  are  made 
at  the  Corliss  Engine  Works,  Providence,  R.I.  A  Wilkinson 
turbine-generator  rated  at  100  kilowatts  for  non-condensing 
service  (six  stages)  is  shown  by  side  and  end  sections  in  Figs.  140 
and  141.  It  will  be  observed  that  in  this  design  the  diaphragms 
are  "dished"  as  in  Curtis  turbines,  while  the  disks  are  flat.  The 
disks  are  made  of  forged  steel,  but  the  blades  are  bronze  castings 
which  are  filed  to  a  sharp  edge  on  the  side  where  the  steam 
enters. 


p.  203 


2O4 


THE  STEAM  TURBINE 


Stage  Packing.  To  prevent  the  leakage  of  steam  between  the 
diaphragms  and  the  shaft  (stage  leakage)^  which  in  some  impulse 
turbines  is  a  considerable  loss  —  often  10  to  20  per  cent.  —  a 
very  ingenious  system  of  steam  packing  has  been  devised.  A 
drawing  illustrating  this  system  is  shown  in  Fig.  142.  By  this 
device,  steam  containing  a  large  amount  of  condensation  is  dis- 
charged into  grooved  packings  between  the  diaphragms  and  the 
shaft  through  ducts  drilled  into  the  hubs  of  the  disks.  This  wet 
steam  is  taken  from  a  part  of  the  labyrinth  packing  at  the  high- 
pressure  end  of  the  turbine  —  through  which  there  is  always 


FIG.  142.     Wilkinson  Labyrinth  Stage  Packing. 

some  leakage  of  steam  —  and  is  conducted  in  the  ducts  shown  in 
the  figure,  which  are  arranged  so  that  the  steam  discharged  into 
a  diaphragm  packing  is  at  a  slightly  higher  pressure  than  that  on 
either  side  of  the  diaphragm.  It  is  probably  possible  in  this  way 
to  practically  eliminate  the  loss  due  to  stage  leakage. 


HAMILTON-HOLZWARTH  TURBINES. 

A  steam  turbine  called  the  Hamilton-Holzwarth  is  being 
developed  by  the  Hooven,  Owens,  Rentschler  Company  of 
Hamilton,  Ohio,  which  is  a  slight  modification  of  the  Rateau 
type.  According  to  designs  which  have  been  published,  this 
turbine  is  divided  into  two  sections  (high-  and  low-pressure) 
which  are  separated  by  a  bearing.  The  principal  difference 


COMMERCIAL  TYPES  •     205 

between  this  and  the  Rateau  type  is  that  the  nozzles  are  arranged 
in  complete  rings  around  the  circumference  of  the  diaphragms  in 
all  the  stages,  instead  of  being  grouped  at  the  high-pressure  end. 
As  in  the  Curtis  turbine,  the  blades  and  nozzles  increase  in  radial 
height  gradually  toward  the  low-pressure  end. 

The  number  of  stages  is  about  the  same  as  in  a  Rateau  turbine 
for  the  same  conditions  of  pressure,  superheat,  and. vacuum,  so 
that  the  nozzles  are  always  designed  to  be  non-expanding.  This 
turbine  has  not  been  developed  commercially,*  so  that  it  is  not 
necessary  to  give  other  details. 


THE   ZOELLY   TURBINE. 

The  Zoelly  turbine  is  a  modified  form  of  the  multi-stage, 
impulse  type.  It  has  fewer  stages  (about  5  to  10),  and  is  gen- 
erally a  much  simpler  design  than  a  Rateau  turbine.  It  repre- 
sents a  noteworthy  attempt  at  increasing  the  steam  velocities  in 
the  blades;  but  with  it  results  the  great  disadvantage  that  the 
surfaces  of  the  numerous  large  wheels  and  blades,  many  of 
which  move  at  high  speeds  in  steam  of  high  pressure,  produce 
excessive  losses  due  to  fluid  friction.  This  fluid  friction  of  disks 
and  blades  increases,  of  course,  enormously  as  the  speed  and  the 
pressure  of  the  steam  are  increased. 

In  a  Zoelly  steam  turbine  there  are  a  number  of  single  impulse 
wheels,  each  rotating  in  a  separate  chamber,  the  walls  of  which 
are  formed  by  stationary  flat  disks  to  which  the  nozzles  are 
attached.  At  the  high-pressure  end  the  nozzles  occupy  only  a 
portion  of  the  periphery;  but  the  area  covered  gradually  in- 
creases till  at  the  low-pressure  end  practically  the  whole  circum- 
ference is  covered.  When  there  are  about  ten  stages  the  pressure 
in  no  stage  drops  to  less  than  .58  of  that  in  the  preceding  stage, 
so  that  non-expanding  nozzles  are  used.  The  blades  are  dove- 
tailed as  represented  by  Fig.  61,  and  there  is  no  shroud  ring.  The 

*  One  of  these  turbines  was  operated  for  a  few  days  at  the  St.  Louis  exposition 
in  1904,  but  the  author  has  heard  of  no  other  important  installation. 


206  THE   STEAM  TURBINE 

tops  of  the  blades  are  cut  off  parallel  to  the  shaft,  but  at  the  roots 
they  are  made  with  a  considerably  greater  height  on  the  dis- 
charge side  than  on  the  entrance  side.  .This  is  of  course  desir- 
able to  allow  for  the  loss  of  velocity  in  the  blades,  but  it  is  stated 
that  the  height  is  made  unusually  large  to  cause  the  steam  to  flow 
smoothly  through  them  without  producing  eddies.*  In  order 
to  accommodate,  in  the  different  stages,  the  size  of  the  nozzles  to 
the  expansion  of  the  steam,  the  radial  widths  of  the  nozzle  parts 
are  gradually  increased  toward  the  low-pressure  end.  The  most 
interesting  part  of  the  design  of  this  turbine  is,  however,  in  the 
construction  of  the  blade  wheels  to  resist  the  stresses  due  to 
extraordinarily  high  peripheral  speeds.  As  the  blades  for  this 
turbine  are  made  at  present,  they -are  much  longer  in  comparison 
with  the  size  of  the  wheel  than  in  any  other  turbine;  in  fact,  the 
length  of  the  blade  is  sometimes  nearly  one-half  the  radius  of  the 
wheel.  These  long  blades  are  tapered  off  toward  the  outer  ends 
in  order  to  make  them  of  uniform  strength.  The  disks  are 
made  of  forged  steel  and  the  blades  of  nickel  steel  which  resists 
erosive  action  very  effectually.  This  simple  construction  of  the 
wheels  and  blades  makes  a  great  saving  in  weight.  The  large 
radial  divergence  of  these  long  blades  makes  possible  the  use  of 
very  small  angles  on  the  discharge  side  of  the  wheel. 

Turbines  of  this  type  intended  for  condensing  service  are 
usually  made  in  two  sections  —  each  about  5  stages  —  placed 
far  enough  apart  to  permit  a  bearing  to  be  located  to  support  the 
turbine  shaft  at  the  middle. 

Zoelly  deserves  the  distinction  of  being  the  first  to  adopt,  in 
'  impulse  turbines,  the  use  of  blades  with  unequal  angles  at  the 
entrance  and  exit  sides.  Simplicity  in  the  design  of  the  working 
parts  is  the  most  striking  feature  of  these  turbines. 

There  are  a  number  of  manufacturers  of  Zoelly  turbines  in 
Germany  and  France.  It  is  stated  that  a  Zoelly  turbine  has  been 
constructed  at  the  Providence  Engineering  Works,  Providence, 
R.I. 

*  It  is  probable  that  there  is  considerable  expansion  of  the  steam  in  such 
blades. 


COMMERCIAL  TYPES  207 

PELTON  AND   SIMILAR  BUCKET  WHEEL  TURBINES. 

Impulse  turbines  with  bucket  wheels  of  the  Pelton  type  have 
recently  received  a  great  deal  of  exploitation  from  inventors  in 
America.  This  type  has  probably  received  so  much  attention 
because  the  Pelton  water  wheel,  commonly  known  as  the  "  hurdy- 
gurdy  "  wheel,  has  proved  so  efficient  in  American  water  power 
plants  where  a  high  head  is  available. 

Professors  Rateau  in  Paris  and  Stumpf  and  Riedler  in  Berlin 
have  done  a  great  deal  of  experimental  work  on  such  turbines, 
but  they  have  practically  abandoned  them  for  those  with  blade 
wheels  of  the  common  axial  flow  type.  Rateau  has  now  adopted 
his  famous  "  multicellular  "  type,  and  Riedler  is  engaged  in 
developing  the  Curtis  turbine  in  Germany. 

Sturtevant  Turbine.  A  steam  turbine  has  been  developed  by 
the  B.  F.  Sturtevant  Company,  Boston,  Mass.,  from  designs 
prepared  by  Mr.  W.  E.  Snow,  which  in  the  general  bucket 
arrangement  is  similar  to  the  old  Riedler-Stumpf  type.*  This 
turbine  was  developed  primarily  for  driving  blowers,  but  it  is, 
of  course,  equally  applicable  for  other  purposes.  It  is  notable 
particularly  for  its  extreme  simplicity  and  strength. 

Fig.  143  is  a  good  illustration  of  this  turbine,  showing  the 
buckets  on  the  wheel  and  the  segments  on  the  inside  of  the 
casing  including  the  nozzles  and  the  stationary  "  reversing  " 
buckets.  Three,  four,  or  five  of  the  latter  are  cut  into  the  seg- 
ment, following  each  nozzle,  depending  on  the  velocity  of  the 
steam.  Fig.  144  shows  more  clearly  the  arrangement  of  the 

*  The  unique  feature  of  the  Riedler-Stumpf  turbine  was  in  the  bucket  wheels, 
of  which  the  Sturtevant  wheel  in  Fig.  143  is  a  good  illustration  except  that  there  was 
usually  a  double  row  of  buckets  on  the  rim  of  each  wheel.  These  wheels  were 
patented  by  Prof.  Stumpf  and  developed  with  the  assistance  of  Prof.  Riedler.  The 
buckets  were  cut  into  the  rim  of  the  wheel  by  a  milling  machine,  and  were  arranged 
to  overlap  each  other  like  the  shingles  of  a  roof,  instead  of  being  placed  one  in  front 
of  another  as  in  a  Pelton  water  wheel.  Unusual  attention  was  given  to  balancing 
the  wheels,  which  were  in  the  form  of  flat  disks.  Stumpf  states  that  these  disks 
were  balanced  so  accurately  that  the  center  of  gravity  came  within  .004  of  the 
diameter  from  the  geometric  center.  These  disks  were  similar  to  the  design  in 
Fig.  216.  It  is  stated  that  such  wheels  were  designed  for  a  factor  of  safety  of  5 
at  a  rim  speed  of  1200  feet  per  second. 


208 


THE   STEAM   TURBINE 


nozzle  with  respect  to  the  stationary  buckets.  The  nozzle  i? 
the  nearly  square  opening  shown  next  to  the  first  bucket,  count- 
ing from  the  left.  As  shown  here  the  steam  flow  will  then  be 


FIG.  143.     Sturtevant  Turbine  with  the  Wheel  Removed  to  the  Side  to  Show  the 
Arrangement  of  the  Buckets. 

toward  the  right  from  the  nozzle  into  the  bucket  opposite  it  on 
the  wheel.  From  this  moving  bucket  the  steam  will  be  diverted 
back  into  the  stationary  bucket  next  to  the  nozzle  and  the  steam 


FIG.  144.     Sturtevant  Nozzle  and  Stationary  Buckets,  showing  Flanged  Connec- 
tion to  the  Steam  Chest. 

path  continues  alternately  through  moving  and  stationary 
buckets  until  the  last  stationary  bucket  has  been  passed,  when 
it  will  escape  into  the  casing  and  into  the  exhaust  pipe.  The 
stationary  bucket  shown  to  the  left  of  the  nozzle  is  called  a 


COMMERCIAL  TYPES 


209 


"  supplementary  "  bucket  intended  to  utilize  the  velocity  of 
the  steam  escaping  over  the  top  of  the  first  moving  bucket  oppo- 
site the  nozzle.  Its  function  is  to  divert  this  steam  leakage  into 
the  moving  buckets. 


FIG.  145.     Section  of  Sturtevant  Turbine. 


Fig.  145  is  a  sectional  view  of  the  turbine.  Hand  wheels  are 
shown  on  valves  by  which  the  flow  of  steam  into  the  nozzles  can 
be  controlled.  It  is  thus  possible  to  close  some  of  the  nozzles 
on  light  loads  and  obtain  nearly  as  good  efficiency  and  steam 
consumption  as  at  full  load.  The  method  is  the  same  as  ex- 
plained for  De  Laval  turbines  on  page  144.  The  governor  of  the 
centrifugal  throttling  type  is  shown  at  the  extreme  right-hand 
end  of  the  turbine  shaft.  It  is  one  of  the  type  with  weights 
acting  on  knife  edges,  in  principle  somewhat  like  the  De  Laval 
governor  (Figs.  1 60  and  161).  Like  other  parts  of  this  turbine 


210 


THE    STEAM   TURBINE 


it  is  made  as  simple  as  possible,  consisting  of  very  few  parts  as 
shown  in  Fig.  147. 

The  main  bearings  have  solid  linings  of  phosphor  bronze. 
They  are  of  the  self-aligning,  ring-oiling  type.  The  weight  on 
these  bearings  never  exceeds  14  pounds  per  square  inch  of  bear- 
ing surface. 

The  speed  of  these  turbines  is  from  1600  to  3000  revolutions 
per  minute.  These  low  speed  limits  compared  with  the  speeds 
of  single-stage  De  Laval  turbines  are  made  possible  by  the 


FIG.  147.     The  Parts  of  a  Sturtevant  Governor. 

application  of  the  velocity  stage  principle  in  the  use  of  the 
reversing  buckets. 

Fig.  148  is  an  illustration  of  a  Sturtevant  turbine  direct-con- 
nected to  a  ventilating  fan  or  blower.  The  governor  mechanism 
is  at  the  left-hand  end.  Valves  for  closing  nozzles  to  adjust  the 
steam  supply  to  the  load,  to  get  the  best  efficiency  of  the  nozzles 
and  blades,  are  shown  clearly  outside  the  casing. 

The  deep  base  which  the  small  diameter  of  this  turbine  neces- 
sitates, is  utilized  for  steam  chambers,  to  which  the  main  admis- 
sion and  exhaust  steam  piping  is  connected.  Overhead  pipes 
axe  in  this  way  eliminated. 


COMMERCIAL  TYPES  211 

The  bucket  wheel  is  a  single  forging  of  open  hearth  steel,  and 
as  the  buckets  are  cut  out  of /the  solid  metal,  a  wheel  of  great 
strength  is  secured.  Blade  breakage  and  "striking"  are  elimi- 
nated, because  if  the  bucket  wheel  should  get  out  of  line  and 
touch  the  casing  on  its  sides,  the  result  would  be  merely  like  the 
rubbing  together  of  two  steel  plates,  which  would  produce  no 
serious  injury. 


FIG.  148.     Sturtevant  Turbine  Direct-connected  to  a  Blower. 

This  turbine  was  designed  to  require  the  minimum  amount 
of  attention  and  repairs.  It  is  stated  that  it  can  be  operated 
continuously  under  ordinary  conditions  with  no  more  attention 
then  the  weekly  filling  of  the  oil  wells  in  the  main  bearings.  It 
is  therefore  particularly  well  suited  for  driving  any  type  of 
auxiliary  machinery,  especially  such  as  may  be  located  in  inac- 
cessible places.  Such  turbines  make  operating  expense  and 
depreciation  low,  and  it  is  stated  by  some  engineers  that  they 
have  operated  turbines  of  this  type  for  five  years  at  a  time 
without  any  expense  for  repairs. 

Kerr  Turbine.  An  impulse  turbine  of  the  Pelton  type  has 
been  patented  by  Mr.  C.  V.  Kerr  and  is  manufactured  by  the 
Kerr  Turbine  Company,  Wellsville,  N.  Y.  In  this  turbine 
typical  Pelton  double  cup-shaped  buckets  are  used  into  which 
jets  of  steam  at  high  velocity  are  discharged  from  nozzles,  located 


>\ 


212 


THE   STEAM  TURBINE 


as  in  Fig.  149,  around  the  periphery  of  the  wheel.     The  inside 
surface  of  each  bucket  is  formed  of  two  intersecting  surfaces  of 


FIG.  149.     Kerr  Bucket  Wheel  and  Nozzles. 


FIG.  150.     Sectional  View  of  the  Kerr  Turbine. 

revolution,  approximately  ellipsoidal,  somewhat  like  the  reflector 
of  a  locomotive  headlight. 

A  section  of  a  Kerr  turbine  is  shown  in  Fig.  150.     In  this 


COMMERCIAL  TYPES 


213 


design  there  are  five  compartments  or  stages,  each  with  a  single 
bucket  wheel.  In  the  design  of  this  turbine  provision  was  made 
for  its  manufacture  in  standarcf  " unit  parts."  In  this  sense  the 
turbine  casing  shown  here  in  section,  consists  of  steam  and  exhaust 
end  castings  and  a  number  of  nozzle  diaphragms  between  the 
ends.  In  the  chambers  thus  formed  steel  disks  revolve,  each 
having  a  row  of  double  buckets  dovetailed  in  the  rim.  By  this 
simple  arrangement  it  is  possible  to  build  up  turbines  of  any 
size  or  for  any  pressure  and  vacuum  by  adding  sections  or  nozzles 
as  may  be  required.  Steam  flowing  from  one  stage  into  the 
next  is  discharged  into  these  buckets  in  tangential  jets  by  the 
nozzles  screwed  into  the  steel  nozzle  bodies,  which  are  accurately 
set  in  place  and  riveted  into  the  diaphragm  castings. 

The  governor  is  of  the  centrifugal  type,  consisting  of  weights 
moving  on  knife-edges.     A  section  of  the  governor  weights  and 


&g^ 


\v         4/^0.  G.  Open  Position 
y*^C.G.  Closed  Position 


Ctr.  Line  of  Shaft  ->" 


Section  A-A 


FIG.  151.     Section  of  Kerr  Governor  Weights  and  Mounting. 

their  mounting  is  illustrated  in  Fig.  151.  The  weights  are  sup- 
ported at  three  points.  The  hardened  knife-edge  at  B  is  straight, 
and  of  sufficient  width  for  the  stresses  on  it.  At  90  degrees  on 
each  side  is  a  rolling  contact  at  C.  The  curve  at  this  point  is 
such  that  the  bearing  between  the  weight  and  the  cam  collar  is 
always  on  the  line  of  centers.  Pure  rolling  contact  is  thus  secured, 
and  the  weight,  without  being  fastened  in  position,  is  firmly 
driven  by  its  triangular  support.  The  outward  movement  of 
the  weights  compresses  the  governor  spring  and  operates,  through 
lever  connections,  a  balanced  piston  valve  controlling  the  flow  oi 
steam. 


214 


THE   STEAM  TURBINE 


Fig.  152  shows  a  typical  Kerr  turbine  direct-connected  to  an 
electric  generator.  Because  of  the  simplicity  of  the  design  these 
turbines  are  particularly  suitable  for  isolated  lighting  plants  and 
for  driving  centrifugal  pumps  and  blowers. 

Terry  Steam  Turbine.  Like  the  Sturtevant  turbine,  the  one 
invented  by  Mr.  Edward  C.  Terry  belongs  to  the  Pelton  impulse 
type  in  which  there  are  two  or  more  velocity  stages.  Stationary 
reversing  buckets  are  arranged  in  groups  —  one  for  each  nozzle  — 
around  the  interior  of  the  casing.  These  bucket  groups  are  shown 
in  Fig.  153,  where  a  Terry  turbine  is  shown  with  the  upper  half 
of  the  casing  raised  for  inspection.  In  this  illustration  there 
are  four  stationary  buckets  for  each  nozzle.  Obviously  the  steam 


I 


FIG.  152.     A  loo-Kilowatt  Kerr  Turbine-Generator. 

is  returned  to  the  moving  buckets  as  many  times  as  there  are 
stationary  buckets  in  each  group.  These  stationary  buckets  are 
made  of  gun  metal,  and  each  has  a  crescent-shaped  hole  at  the 
center  through  which  the  steam  partially  exhausts.  There  is, 
therefore,  apparently  considerable  expansion  in  the  moving 


COMMERCIAL  TYPES 


215 


blades.     A  valve  is  provided  for  each  nozzle,  so  that  when  it  is 
desired  some  of  them  can  be  closed.     Speeds  of  these  turbines 


FIG.  153.     Terry  Turbine  with  the  Casing  Raised. 

vary  from  2500  for  a  ic-horsepower  size  to  1600  for  3oo-horse- 
power. 

Fig.  154  is  an  illustration  of  a  Terry  turbine  direct-connected  to 
a  five-stage  high-pressure  turbine  pump. 


216 


THE   STEAM  TURBINE 


COMMERCIAL  TYPES 


217 


Dake  Steam  Turbine.  This  turbine  has  stepped  buckets,  but 
the  nozzles  do  not  discharge  radially.  Stationary  and  moving 
buckets,  with  a  section  of  the  bucket  wheel,  are  shown  in  Fig.  155. 


Steam  Port 
Expansion  Wedge 
Nozzle 


Bucket 
Guide  Passage 


FIG.     156.       Diagram     of     Nozzles     and 
FIG.  155.     Steam  Passages  in  a  Dake         Buckets  of  a    Dake   Turbine    Showing 
Turbine.  Expansion  Wedges. 

This  arrangement  is  unique  in  that  the  steam  passes  through  the 
stationary  buckets  b  in  an  axial  direction,  and  is  deflected  radi- 
ally by  the  two  sets  of  buckets  a  and  c  on  the  wheel.  Steam 
enters  the  nozzles  from  the  steam  chest  S.  Relative  positions  of 
nozzles  and  buckets  are  illustrated  diagrammatically  in  Fig.  156. 
Actually  in  the  turbine  the  horizontal  lines  in  this  figure  are,  of 
course,  arcs  of  circles.  The  "steps"  shown  here  in  the  wall  of 
the  bucket  ring  are  intended  to  bring  the  surfaces  upon  which 
the  steam  impinges  nearer  to  the  nozzles  and  to  present  always 
approximately  the  same  angle  to  the  flow  of  steam. 

The  nozzles  are  designed  with  the  object  of  delivering  the  steam 
to  the  buckets  in  parallel  jets.  Throughout  their  lengths  the 
nozzle  walls  are  the  same  distance  apart,  and  expansion  is  secured 
by  the  use  of  "  expansion  wedges,"  shown  plainly  in  both  figures, 
which  are  set  centrally  in  the  nozzles.  These  wedges  can  be 
readily  removed  and  replaced,  so  that  it  is  not  difficult  to  insert 
a  wedge  properly  proportioned  to  give  the  best  expansion  for  a 
given  steam  pressure. 


THE    STEAM   TURBINE 


p 

£ 
a 


COMMERCIAL   TYPES 


SPIRO  TURBINES. 

The  Spiro  turbine  (Fig.  157)  consists  simply  of  two  herringbone 
gear  wheels  which  mesh  together  and  revolve  in  a  close-fitting 
casing.  Steam  enters  at  the  inlet  pipe  at  the  bottom  and  passes 
around  the  gears  in  its  expansion  to  the  exhaust  pipe  at  the  top. 
Steam  discharges  from  the  inlet  pipe  through  the  small  holes, 
equivalent  to  non-expanding  nozzles  shown  in  Figs.  158  and  159, 


FIG.  158.     The  Spiro  Casing  or  Cylinder.    The  two  holes  near  the  central  rib 
inside  the  cylinder  are  the  steam  nozzles. 

into  the  "  pockets  "  or  spaces  between  adjacent  gear  teeth.  As 
the  rotors  revolve  the  "  tooth-space  "  occupied  by  the  steam 
increases  in  length  as  the  steam  expands.  Finally  the  steam 
escapes  when  the  outer  ends  of  the  teeth  pass  the  line  of  contact 
between  the  two  rotors.  The  increased  length  of  this  "  tooth- 
space  "  from  the  time  the  steam  is  admitted  until  it  is  exhausted 
is  shown  in  Fig.  160,  by  the  comparison  of  the  length  of  the 
tooth-grooves  at  "A"  with  the  length  of  the  outer  white  lines. 
By  having  the  steam  inlet  at  the  bottom  the  weight  of  the  rotors 
is  partly  carried  by  the  steam  pressure  and  friction  is  much 
reduced  below  what  it  would  be  if  the  inlet  were  at  the  top. 


THE   STEAM   TURBINE 


Spiro  turbines  are  suitable  only  for  non-condensing  operation 
and  find  application  usually  for  driving  the  auxiliaries  like 
blowers  and  pumps  in  a  power  plant  or  in  office  buildings  where 
the  exhaust  steam  is  passed  through  feed- water  heaters  or  is  used 


EXHAUST 


INLET- 
FIG.  159.     Section  of  Spiro  Cylinder  and  Rotors  at  Mid-length. 


FIG.  160.    The  Rotors. 

for  heating  buildings.  Under  these  conditions  low  steam  con- 
sumption, as  would  be  obtainable  with  condensing  operation,  is 
unimportant.  Sufficient  expansion  for  condensing  operation  is 
impracticable  with  this  type  of  turbine.  Compactness  is  also 
an  important  feature.  The  casing  of  such  a  turbine  is  alone  no 


COMMERCIAL  TYPES 

larger  than  the  cylinder  of  a  good  high-speed  engine  of  the  same 
capacity  and  is  even  smaller  than  comparable  commercial  sizes 
of  electric  motors.  Governing  is  accomplished  by  throttling  the 
steam  pressure  by  a  method  similar  to  that  used  for  nearly  all 
steam  turbines  of  relatively  small  size  (less  than  100  horsepower). 


CHAPTER    VIII. 
GOVERNING  STEAM  TURBINES. 

METHODS  of  governing  steam  turbines,  or,  in  other  words,  of 
regulating  the  supply  of  steam  to  suit  the  load  on  the  machine, 
may  be  classified  as  follows: 

1.  Throttling  or  partly  closing  the  steam  admission  valve. 

2.  Varying  the  cross-section  of  the  steam  passages  by  "  cutting 
out  nozzles." 

3.  Varying  the  time  of  admission,  or  "blast"  governing. 

4.  Admitting  steam  at  boiler  pressure  at  various  points  along 
the  direction  of  steam  flow,  or  " by-pass"  governing. 

When  the  steam  admission  of  a  turbine  is  partly  closed,  the 
amount  of  steam  passing  through  the  valve  ports  is,  of  course, 
reduced;  but  at  the  same  time  the  steam  is  throttled,  meaning 
that  the  pressure  is  reduced  without  changing  the  heat  contents 
in  a  unit  weight.  Although  in  this  throttling  process  the  total 
heat  in  a  pound  of  steam  remains  unchanged,  the  energy  avail- 
able from  expansion  is  considerably  reduced.  If  steam  at  165 
pounds  per  square  inch  absolute  pressure  which  contained  ini- 
tially 2  per  cent,  moisture  is  throttled  without  loss  of  heat,  that 
is,  without  doing  work,  to  25  pounds  per  square  inch  absolute, 
the  steam  at  this  lower  pressure  will  have  43  degrees  F.  of  super- 
heat. Now  if  the  available  energy  is  calculated  for  adiabatic 
expansion  from  this  lower  pressure  and  43  degrees  F.  super- 
heat to  i  pound  absolute,  it  is  found  to  be  207  B.T.U.  The 
available  energy  for  adiabatic  expansion  of  steam  at  the  initial 
condition  before  throttling  (containing  2  per  cent,  moisture) 
to  the  same  final  pressure  is,  on  the  other  hand,  316  B.T.U.* 

*  Although  the  moisture  is  removed  and  the  steam  is  superheated  there  is  no 
gain  to  offset  the  loss  in  available  energy  except  that  the  disk  and  blade  rotation 
losses  are  reduced;  but  the  gain  from  this  cause  could  not  probably  exceed  10  per 
cent.  "  Drying  "  action  then  is  not  very  important,  and  it  will  have  very  little 
influence  in  remedying  large  losses  due  to  throttling. 

218 


GOVERNING  STEAM  TURBINES 


2I9 


In  this  extreme  case  of  throttling,  the  available  energy  of  the 
steam  is  reduced  about  35  per  cent,  and  consequently,  for  the 
same  work,  approximately  35  percent,  more  steam  is  needed  with 
throttling  valves  than  if  the  steam  could  be  used  at  light  loads 
without  throttling.  It  is  not  unusual  for  turbines  governed  by 
throttling  to  take  steam  at  full  load  at  135  pounds  pressure 
when  the  steam  supplied  is  at  165  pounds.  Then  the 
maximum  pressure  becomes  available  only  on  overload  just  before 
the  stage  valves  open.  Efficiency  of  an  expanding  nozzle  is 
considerably  reduced  when  it  is  used  with  pressures  very  much 
different  from  that  for  which  it  was  designed,  as  shown  by  the 
curve  in  Fig.  28.  Blade  efficiencies  are  similarly  reduced  when 
the  available  energies  and  consequently  the  velocities  are  not 
those  for  which  the  blades  were  designed.  Fig.  161  shows  very 


FIG.  161.     Effect  of  Throttling  on  Steam  Consumption. 

plainly  the  effect  of  throttling  on  the  economy  of  steam  turbines. 
The  two  curves  in  this  figure  show  the  steam  consumption  per 
electrical  kilowatt  for  a  6oo-kilowatt  turbine  when  operating 
(i)  with  a  throttling  governor,  and  (2)  with  a  governor  varying 
the  steam  supply  by  changing  the  area  of  the  steam  passages,  that 
is,  governing  without  appreciable  throttling.  In  spite  of  these 
defects,  however,  governing  by  throttling  has  been  fairly  satis- 
factory. 

In  the  De  Laval,  Rateau,  and  Zoelly  turbines  governing  is 
effected  by  throttling  devices.  For  most  turbines,  the  governor 
itself  is  similar  to  centrifugal  governors  used  in  reciprocating 
engine  practice. 


220 


THE   STEAM  TURBINE 


FIG.  162.     De  Laval  Governor  and  Vacuum  Valve. 


FPG.  163.     Section  of  the  Main  Admission  Valve  of  a  De  Laval  Turbine. 


GOVERNING  STEAM  TURBINES  221 

Fig.  162  shows  cross-sections  of  a  typical  De  Laval  governor. 
It  consists  of  two  half  cylinders  B,  B  which  are  pivoted  in'  a  short 
outer  casing  by  the  knife-edge  A.  Inside  the  casing  these  cylin- 
ders are  fitted  with  pins  C,  C  which  press  on  a  collar  D  when  the 
other  ends  of  these  cylinders  (at  B)  are  thrown  out,  or  tend  to 
separate,  by  centrifugal  force.  The  pressure  on  the  collar  D 
transmitted  by  the  pins  compresses  the  springs  and  forces  a 
central  spindle  G  toward  the  right,  which  moves  with  it  the 
bell-crank  L.  This  bell-crank  moves  a  short  shaft  which  passes 
through  the  steam  pipe  and  has  attached  to  its  other  end  by 
means  of  a  set-screw  a  lever  (shown  in  the  section  of  the  valve, 
Fig.  163)  operating  the  main  admission  valve.  The  weight  of 
the  valve  and  levers  is  balanced  by  the  small  spring  N.  The 
bell-crank  L  has  a  certain  "play"  in  M  which  is  adjusted  to 
make  the  governor  not  too  sensitive  to  momentary  changes  in 
speed.  The  valve  travel  is  only  about  one-eighth  of  an  inch 
from  the  closed  to  the  wide  open  position. 

The  governor  frame  is  supported  on  a  tapering  rod  E  which  is 
fitted  into  the  end  of  the  main  turbine  shaft  K. 

With  condensing  De  Laval  turbines  a  vacuum  valve  T  is 
arranged  in  connection  with  the  governor  to  act  as  an  emergency 
stop  valve.  In  case  the  turbine  exceeds  the  allowable  speed 
limit  due  to  the  failure  of  the  main  admission  valve  to  operate 
properly,  the  vacuum  valve  admits  air  to  the  turbine  exhaust 
pipe  through  the  passage  P.  The  steam  consumption  when 
operating  non-condensing  is  so  much  greater  than  when  condens- 
ing that  it  is  said  to  be  impossible  to  exceed  the  rated  speed  when 
exhausting  into  the  atmosphere,  with  all  the  nozzles  and  the  main 
admission  valve  open. 


FIG.  164.     A  Slide  Valve  Arrangement  for  a  Turbine. 

Governing  by  "  Cutting  Out  Nozzles."     One  of  the  simplest 
forms  of  governing  is  represented  in  Fig.  164  showing  a  plain 


222  THE  STEAM  TURBINE 

slide  valve  arrangement  for  regulating  the  flow  of  steam  through 
a  series  of  nozzles.  This  is  one  of  the  best  systems  that  can  be 
employed  for  an  impulse  turbine  if  an  elaborate  valve  gear  is  to 
be  avoided. 

As  the  full  initial  pressure  is  always  maintained  in  all  the  nozzles 
that  are  open,  there  can  be  very  little  throttling  except  when  the 
valve  is  in  a  position  so  that  one  of  the  nozzles  is  partly  covered. 
The  loss,  however,  due  to  this  amount  of  throttling  is  practically 
negligible  for  other  than  very  light  loads.  Valve  gears  have  been 
designed  to  improve  on  this  slide  valve  method  by  providing  a 
separate  valve  (usually  of  the  poppet  type)  for  each  nozzle  or, 
at  most,  for  a  small  group  of  nozzles.  These  valves  are  opened 
and  closed  suddenly  by  the  governing  apparatus  by  the  use  of 
either  springs  or  dash-pots,  very  much  as  with  our  modern  Cor- 
liss valve  gears  for  reciprocating  engines.  The  difficulty  with 
this  last  method  is,  however,  that  there  will  be  abrupt,  although 
perhaps  small,  variations  in  speed  every  time  a  valve  opens  or 
closes,  unless  special  precautions  are  taken  in  the  design.  If 
the  service  is  for  electric  lighting,  speed  irregularities  due  to  such 
governing  may  be  sufficient  to  produce  a  flicker  in  the  lights. 

In  turbines  with  more  than  one  pressure  stage,  as,  for  example, 
in  the  Curtis  and  Rateau  types,  it  has  often  been  proposed  to 
control  the  admission  to  each  stage.  Apparently  the  only  objec- 
tion to  such  a  scheme  would  be  in  the  very  complicated  valve 
gear  that  would  be  needed;  but,  contrary  to  what  one  might 
expect,  it  can  be  shown  by  tests  and  demonstrated  mathemati- 
cally that  such  an  arrangement  would  not  give  as  good  economy 
as  if  only  the  first  stage  is  controlled. 

The  only  advantage  resulting  from  this  method  of  controlling 
the  steam  supply  is  that  by  making  the  light  load  pressures  more 
nearly  in  the  same  proportion  to  each  other  as  for  full  load  and 
overload,  the  stresses  in  the  diaphragms  separating  the  stages  are 
more  nearly  the  same  as  calculated  in  the  original  design.  There 
is  probably  no  commercial  type  of  turbine  using  such  a  compli- 
cated method  of  governing  except  for  large  overloads,  when  econ- 
omy is  not  of  importance  and  the  conditions  are  more  of  emergency 


GOVERNING   STEAM  TURBINES  223 

than  of  continuous  operation.  Governors  for  the  larger  sizes  of 
the  Curtis  turbines  show  the  merits  of  this  method  to  the  best 
advantage.  By  governing  in  this  way  it  is  possible  to  vary  the 
number  of  valves  supplying  steam  to  the  turbine  in  proportion 
to  the  size  of  the  load,  thus  maintaining  a  constant  initial  pres- 
sure and  therefore  constant  velocity  in  the  nozzles  and  blades. 
In  a  single  stage  turbine  there  is  no  difficulty  in  applying  this 
method,  and  consequently  the  energy  and  velocity  are  always 
those  suited  for  the  best  nozzle  and  blade  efficiency.  Usually 
in  a  turbine  of  several  stages  no  attempt  is  made  to  regulate 
the  number  of  nozzles  after  the  first  stage,  on  account  of  the 
mechanical  difficulties  inseparable  from  a  complicated  valve 
gear.  In  order  to  secure  a  correct  energy  and  velocity  distri- 
bution throughout  the  turbine,  the  nozzles  in  all  the  different 
stages  should,  of  course,  be  changed  in  the  same  ratio.  This 
scheme  is  not  impossible  and  has  been  attempted  in  some  German 
designs.  With  turbines  like  the  Rateau  and  Parsons,  where  the 
drop  of  pressure  is  very  small  in  each  stage,  and  where  there 
are,  therefore,  a  great  many  stages,  any  method  of  cutting  out 
some  of  the  steam  passages  to  reduce  the  area  at  Jight  loads  is 
impracticable. 

Types  of  governors  to  be  used  depend  a  great  deal  on  the 
capacity  and  the  kind  of  service.  The  smaller  sizes  have  usually 
simple  forms,  while  the  larger  ones  are  necessarily  more  com- 
plicated. On  the  small  turbines,  where  an  elaborate  valve  gear 
is  not  desirable,  the  valves  are  moved  by  the  direct  action  of  the 
centrifugal  force  of  the  governor  weights.  This  is  called  direct 
governing,  to  distinguish  it  from  the  " relay"  system  used  by 
most  turbine  manufacturers  for  large  machines.  By  the  direct 
method  a  comparatively  large  centrifugal  force  is  necessary  to 
move  the  valves;  and  unless  they  are  carefully  balanced  it  is 
difficult  to  make  the  governor  sensitive  to  fluctuations  in  the 
load.  Besides,  if  for  any  reason  a  valve  sticks,  there  may  be 
wide  variations  in  speed. 

By  the  indirect  or  what  is  commonly  called  the  "relay" 
method  the  centrifugal  force  of  the  governor  is  needed  only  to 


224 


THE   STEAM  TURBINE 


226  THE   STEAM  TURBINE 

"give  the  signal,"  as  we  may  say,  which  sets  in  motion  an  auxil- 
iary mechanism  by  which  the  valves  are  moved  by  gearing  con- 
nected to  the  main  shaft  or  by  steam  or  hydraulic  pressure.  In 
Curtis  turbines  of  all  sizes  up  to  500  kilowatts  the  valves  are 
operated  mechanically,  and  for  larger  sizes  a  hydraulic  apparatus* 
is  used. 

Electromagnetic  Control  of  Valves.  Formerly,  in  large  Curtis 
turbines,  the  valves  opening  the  nozzles  were  operated  by  the 
pressure  of  steam  admitted  through  a  port  opened  and  closed  by 
a  " pilot"  valve  controlled  by  electromagnets.  The  governor 
was  connected  to  a  very  simple  mechanism  for  the  purpose  of 
making  and  breaking  the  current  through  the  electromagnets, 
which,  in  turn,  moved  the  "pilot"  valves  operating  the  main 
valves  on  the  turbine. 

Mechanical  Valve  Control.  One  of  the  recent  developments 
in  the  valve  gears  for  large  turbines  governed  by  cutting  out 
nozzles  is  the  successful  replacing  of  the  electromagnetic  "relay" 
outfit  formerly  used  on  Curtis  turbines  by  a  positive  mechanical 
valve  gear,  due  to  Mr.  Richard  H.  Rice. 

This  valve  mechanism  is  well  illustrated  in  Figs.  165  and  166, 
where  it  is  shown  applied  for  regulating  the  steam  admission  to 
the  first  stage  of  an  impulse  turbine.  Steam  in  the  steam  chest 
C  is  maintained  constant  at  the  pressure  for  which  the  turbine 
was  designed,  and  the  valves  are  operated  so  that  they  are  always 
wide  open  or  else  tightly  closed.  When  the  valve  rod  t,  Fig.  166, 
is  raised  steam  is  admitted  through  the  port  A,  from  which  it 
passes  into  a  nozzle  plate  (like  Fig.  114)  at  B  to  be  discharged  at 
high  velocity  into  the  blades  of  the  first  stage. 

The  valve  gear  consists  essentially,  besides  the  worm  gears 
shown  at  the  right-hand  side  of  the  figures,  of  a  connecting  rod 
moving  a  bell-crank  1,  to  which  two  dogs  or  " catches,"  w,  w,  are 
attached  by  pins.  The  extreme  ends  of  these  dogs,  marked  i 
will  engage  with  the  teeth  on  the  steel  plates  u  and  v.  An 
•eccentric,  h  (Fig.  165),  gives  the  connecting  rod  k  a  reciprocat- 
ing motion  which,  being  transmitted  to  1,  moves  the  dogs  w,  w 
up  and  down.  In  Fig.  166  the  lower  dog  is  shown  sliding  on 


GOVERNING   STEAM  TURBINES  227 

the  plate  u,  and  in  its  lowest  position  it  touches  the  tooth  on 
this  plate.  The  upper  dog  is  kept  out  of  contact  with  the  tooth 
on  the  plate  v  by  the  lever  x,  which  by  engaging  with  the  lower 
end  of  this  dog,  marked  2  in  the  figure,  raises  the  end  i  out  of 
reach  of  the  tooth.  The  letters  x  and  s  are  at  opposite  ends  of 
the  same  lever  supported  on  the  shaft  m.  In  the  top  view  of 
this  valve  gear  shown  in  Fig.  165  there  are  five  valves  operated 
by  the  connecting  rod  k.  On  the  same  eccentric,  h,  there  is 
also  another  similar  connecting  rod,  j,  operating  five  valves  on 
the  opposite  side  of  the  turbine.  The  steam  supply  of  the  tur- 
bine is  therefore  regulated  by  ten  valves. 

The  position  of  the  end  of  the  lever  at  x  is  regulated  by  means 
of  the  rod  q,  which  is  connected  to  the  Curtis  governor  illustrated 
in  Fig.  167.  Speed  regulation  by  means  of  this  governor  is 
accomplished  by  the  balance  maintained  between  the  centrifugal 
effort  of  moving  weights  and  the  static  forces  exerted  by  springs. 
The  governor  is  keyed  to  the  main  turbine  shaft  at  S  and,  of 
course,  rotates  with  it.  It  is  protected  on  two  sides  by  a  stationary 
looped  casing,  of  which  a  section  is  shown  at  the  top  of  the  figure. 
In  the  order  of  action  of  this  governor  the  weights  A  fly  out  on 
account  of  centrifugal  force,  moving  on  knife-edges  near  their 
largest  diameter,  and  pull  down  the  governor  rod  C  by  the  pres- 
sure exerted  on  other  smaller  knife-edges  B.  The  governor  rod 
is  pulled  down  against  the  action  of  the  heavy  spring  D.  At  E 
a  ball-bearing  gimbel  joint,  thoroughly  lubricated,  forms  a 
junction  point  between  the  revolving  shaft  of  the  turbine  and 
the  stationary  lever  of  the  governor  (shown  in  the  figure 
extending  toward  the  right,  nearly  horizontally).*  This 
stationary  lever  is  connected  by  means  of  a  bell-crank  to  the 
rod  q  (Fig.  166)  and  thus  determines  the  position  of  the 
lever  x. 

To  illustrate  the  action  of  this  valve  gear  and  the  governor, 
assume  the  load  on  the  turbine  has  been  increased  and  the  speed 

*  Connected  to  the  stationary  lever  of  the  governor  is  an  auxiliary  spring  F  fer 
varying  the  speed  when  synchronizing.  By  means  of  a  small  motor  G  the  tension  of 
this  spring  can  be  adjusted  from  the  switchboard. 


228 


THE   STEAM  TURBINE 


has  dropped  a  little,  indicating  that  more  steam  is  needed  and 
that  the  valves  have  been  so  arranged*  that  the  one  shown  in 
Fig.  1 66  is  the  next  to  be  opened.  With  a  reduced  speed  the 
governor  weights  A  (Fig.  167)  will  move  in  slightly  toward  the 
center,  reducing  the  tension  on  the  governor  spring  D,  so  that 
the  rod  C  and  the  left-hand  end  of  the  stationary  governor  lever 
are  raised.  By  means  of  an  auxiliary  lever  and  a  bell-crank  the 
rod  q  is  raised  and  the  end  x  of  the  lever  attached  to  it  is  lowered 


FIG.  167.     Sectional  View  of  Curtis  Governor. 

to  engage  the  catch  2  of  the  lower  dog  w,  releasing  at  the  same 
time  the  upper  dog,  which  now  comes  into  contact  with  the  tooth 
in  the  plate  v,  raises  the  valve  rod  t,  and  admits  steam  through 

*  Each  of  the  levers  s  controlling  the  dogs  is  set  at  a  little  different  angle 
to  the  horizontal.  The  lever  which  has  its  end  (x)  lowest  will  open  its  valve 
first. 


GOVERNING  STEAM  TURBINES  229 

the  port  A.  When  again  the  speed  becomes  too  high  the  rod  q 
is  lowered,  x  is  raised,  and  the  lower  dog  closes  the  valve.  The 
dogs  are  held  in  position  when  not  in  contact  with  the  lever  x 
by  the  flat  springs  o.  • 

The  eccentric  h  is  moved  by  means  of  gears  connected 
to  the  main  turbine  shaft  S.  A  ring  on  this  shaft  has  a 
single  tooth  a,  which  engages  with  a  gear  wheel  b,  on  the 
shaft  c,  which  by  means  of  worm  gearing  is  connected  with 
another  horizontal  shaft  f,  and  thus  moves  the  eccentric  h. 
The  hub  of  a  turbine  wheel  is  shown  at  H  (Fig.  166),  and 
carbon  packing  rings  to  prevent  the  leakage  of  steam  from  the 
first  stage  are  illustrated  by  D  and  E.  The  speed  reduction  is 
designed  sometimes  for  one  worm  gear  instead  of  two  as  shown 
here. 

Hydraulic  Motor  Control  of  Valves.  The  hydraulic  governing 
device  used  in  the  designs  of  Curtis  turbines  of  the  500  to 
9000  kilowatts  sizes  is  illustrated  in  the  following  figures.  The 
movement  of  the  horizontal  governor  lever  shown  in  Fig.  167 
is  transmitted  through  the  rod  D  (Fig.  168)  to  a  second  lever 
arm  C  operating  the  pilot  valve  of  the  oil  cylinder  B.  The  piston  A 
operates  the  main  power  arm  of  the  mechanism,  which  trans- 
mits the  motion  either  by  a  rack  connecting  with  a  pinion,  or  by 
means  of  cranks,  to  the  "side"  rod,  shown  in  Fig.  169,  carrying 
the  cams  for  operating  the  valves.  These  cams  act  directly  on 
the  valves,  opening  and  closing  them  according  to  the  demands 
of  the  load.  Because  this  device  has  a  very  slow  motion  it  has 
the  advantage  of  being  practically  independent  of  lubrication  for 
its  successful  operation. 

Governing  by  Varying  the  Time  of  Admission.  Governing 
by  periodic  admission  or  by  "blasts"  was  invented  by  Parsons 
and  has  been  applied  to  practically  all  types  of  steam  turbines 
using  his  name.  In  its  ideal  form  steam  is  admitted  to  the  turbine 
by  a  poppet  valve  in  puffs  or  blasts  in  periods  of  long  or  short 
duration  depending^  on  the  demands  of  the  load.  The  method 
is  explained  usually  by  saying  that  there  are  alternate  periods 
when  the  turbine  casing  is  either  filled  with  steam  or  there  is  no 


230 


THE   STEAM  TURBINE 


FIG.  1 68.     The  Hydraulic  Operating  Mechanism  for  Valves  of  a  Curtis  Turbine. 

steam  at  all.  At  light  loads  the  valve  opens  for  short  periods, 
remaining  closed  the  greater  part  of  the  time.  When  the  load 
increases  the  valve  remains  open  longer,  and  at  about  full  load 


GOVERNING   STEAM  TURBINES 


231 


there  is  full  pressure  in  the  high-pressure  blades,  the  valve 
merely  vibrating  without  sensibly  affecting  the  pressure  of  the 
steam  in  the  passages.  It  is  thought  that  in  this  way  the  full 
benefit  of  high-pressure  steam  can  be  secured  at  all  loads.  This 


fS 
1? 

C/3  £ 


U 


II 

H  8 

u    i! 


is    the    ideal    condition,    but    practical    considerations    greatly 
modify  it. 

Brown-Boveri-Parsons  Governing  Device.  The  method  of  reg- 
ulating the  steam  supply  by  intermittent  admissions  or  "blasts" 
is  typical  of  nearly  all  the  governing  devices  fitted  to  Parsons 


232 


THE   STEAM  TURBINE 


turbines.  The  design  used  for  the  Brown-Boveri-Parsons  tur- 
bines is  illustrated  by  Fig.  171.  Steam  enters  the  turbine  through 
a  main  admission  valve  N  which  is  given  a  vertically  oscillating 
motion.  A  small  piston  mounted  above  this  valve  and  on  the 
same  spindle,  has  steam  at  the  pressure  in  the  main  steam  pipe 
on  its  lower  face  acting  against  the  pressure  of  a  strong  spring 
on  its  upper  face.  An  auxiliary  valve  fitted  on  the  spindle  L  is 
given  an  oscillating  motion  by  an  eccentric  on  the  governor  shaft 
at  M,  which  causes,  at  every  stroke,  the  small  passage  at  the 


FIG.   171.     Brown-Boveri-Parsons  Governing  Device. 

lower  face  of  the  piston  to  communicate  with  the  exhaust, 
making  the  main  valve  N  fall  upon  its  seat.  The  spindle  L  is 
linked  up  to  a  collar  sliding  on  the  governor  shaft.  The  height 
of  the  governor  balls  determines  the  position  of  this  collar.  Thus 
the  height  of  the  governor  augments  or  diminishes  the  amplitude 
of  the  oscillations  of  the  auxiliary  valve  on  L,  and  in  consequence 
causes  the  main  valve  N  to  open  a  longer  or  a  shorter  time  at 
each  admission  of  steam.  The  frequency  of  the  steam  admis- 
sions is  about  150  to  250  per  minute  according  to  the  speed  of 
the  turbine. 


GOVERNING  STEAM  TURBINES 


233 


Westinghouse-Parsons  Governor  and  Valve  Gear.  Diagram- 
matically,  the  governor  and  valve  gear  of  a  Westinghouse- 
Parsons  turbine  are  shown  in  Fig.  172.  A  small  pilot  valve, 
marked  A  in  the  figure,  is  actuated  directly  by  the  governor  by 
means  of  levers  and  links.  This  pilot  valve  controls  the  steam 
supply  of  the  turbine  by  regulating  the  operation  of  the  main 
poppet  admission  valve  which  opens  and  closes  at .  uniform 
intervals  when  the  turbine  is  in  operation.  Speed  variations 


FIG.  172.     Diagrammatic  Arrangement  of  the  Governing  Mechanism  of  a 
Westinghouse-Parsons  Turbine. 

change  the  height  of  the  governor  balls  which,  in  turn,  change 
the  position  of  the  collar  F  of  the  lever  on  the  governor  spindle. 
By  means  of  a  system  of  links  this  lever  varies  the  throw  of  the 
pilot  valve  relatively  to  the  valve  port.  This  pilot  valve  controls 
the  main  admission  valve  by  means  of  the  auxiliary  piston  valve  B 
in  the  same  way  as  in  the  Brown-Boveri  design  which  has 
already  been  explained.  Reciprocating  motion  for  operating 
the  valve  mechanism  originates  in  an  eccentric  driven  by  the 


234  THE   STEAM  TURBINE 

turbine  from  a  worm  on  the  main  shaft.  This  eccentric  gives  an 
oscillating  motion  to  the  levers  supported  at  D,  F,  and  E. 

The  governor  is  of  the  fly-ball  type,  the  ball  levers  being 
mounted  on  knife-edges  instead  of  pins,  to  secure  sensitiveness. 
The  speed  of  the  turbine  may  be  varied,  while  running,  within 
the  limits  of  the  governor  spring  by  grasping  a  knurled  hand 
wheel  at  the  top  of  the  governor  and  bringing  the  spring  and 
tension  nuts  to  rest.  Adjustment  of  the  tension  of  the  spring 
can  then  be  made.  This  device  is  particularly  useful  for  syn- 
chronizing the  speed  of  small  turbine-alternators  operating  in 
parallel,  or  for  distributing  the  load  between  them.  For  syn- 
chronizing large  Westinghouse  turbine-alternator  units  a  small 
motor  controlled  from  the  switchboard  is  used  to  adjust  the 
governor  spring. 

Allis-Chalmers  Governor  Mechanism.  The  governing  device 
of  the  Allis-Chalmers  steam  turbine  is  of  the  Parsons  type,  using 
hydraulic  instead  of  steam  pressure.  The  governor  is  required 
to  operate  a  small  balanced  oil  relay  valve  only,  while  the  two 
steam  valves,  main  and  by-pass,  are  controlled  by  oil  pressures 
of  about  20  pounds  per  square  inch,  acting  upon  a  piston  of 
suitable  size..  The  by-pass  valve  opens  when  the  turbine  is 
required  to  develop  overload  or  the  vacuum  fails. 

The  oil  supply  to  the  bearings  and  to  the  governor  can  be 
interconnected  so  that  the  governor  will  shut  off  the  steam  if  the 
oil  supply  fails. 

"  Blast "  Governing  Compared  with  Throttling.  When  the 
main  steam  admission  valve  of  a  Parsons  turbine  closes  there  is 
still  some  steam  in  the  turbine  casing,  and  this  steam  expands, 
of  course,  to  fill  the  space.  The  same  effect  occurs  also  when 
the  valves  are  first  opened,  and  the  steam  rushes  into  a  region 
of  very  low  pressure.  In  these  two  ways  low  pressures  are  pro- 
duced just  as  with  throttling  valves,  although,  for  the  same 
average  pressure,  the  loss  is  not  nearly  so  great.  The  pressure 
variation  for  a  i5oo-kilowatt  Parsons  turbine  at  one-quarter, 
three-quarter,  and  full  load  is  shown  in  Fig.  173.  At  a 
little  overload  there  is  practically  no  variation  because  the 


GOVERNING   STEAM  TURBINES 


235 


steam  valve  is  then  closed  for  shorter  periods.  Probably  the 
greatest  disadvantage  from  this  method  of  governing  results 
from  " initial  condensation  at  light  loads."  There  is  usually  one 
blast  or  puff  of  steam  in  every  thirty  revolutions.  Steam  admis- 
sions are,  therefore,  far  enough  apart  to  allow  the  interior  of 
the  turbine  to  be  cooled  by  the  falling  temperature  between  the 
blasts.  Now  when  there  is  a  fresh  admission  the  steam  comes 
into  contact  with  the  relatively  cooler  walls  of  the  interior  of  the 


—2570 


—2240 


1270 


Atmospheric  =0 

Absolute  Zero 

Ilodgkinson,  F. 
FIG.  173.     Indicator  Cards  Showing  Initial  Pressures  in  a  Parsons  Steam  Turbine. 

turbine,  and  condensation  must  take  place  just  as  in  a  recipro- 
cating engine. 

If  it  were  possible,  practically,  the  number  of  "periods"  would 
be  made  so  small  that  free  expansion  would  be  reduced  to  a 
minimum;  but  for  a  satisfactory  speed  regulation  long  periods 
are  not  permissible.  It  appears,  therefore,  that  unless  the 
periodicity  can  be  made  low,  the  economy  at  light  loads  is  no 
great  improvement  on  the  method  of  plain  throttling.  A  very 
important  feature  of  this  method,  however,  should  not  be  over- 
looked. This  is  the  advantage  of  having  a  valve  mechanism 
which  is  constantly  moving,  precluding  the  possibility  of  "sticky" 
valves. 


236  THE  STEAM  TURBINE 

The  time  required  for  the  steam  entrapped  in  the  casing  when 
the  valves  are  closed  to  drop  in  pressure  by  a  given  amount  can 
be  calculated  very  simply  as  follows. 
Let 

Wl  =  weight  of  steam   (pounds)  entrapped  when  the  valves 

are  closed, 
W    =  weight  of   steam   (pounds)  in  the  turbine  casing  after 

expanding  a  time  /, 
W  =  weight   of   steam    (pounds)   flowing   per  second   when 

the  valves  are  wide  open,  that  is,  when  the  pressures 

in   the  casing   are   those    for  which  the  blading  was 

designed, 
Pl  =  initial  absolute  pressure  of  the  steam  delivered  to  the 

turbine, 
P2  =  final  absolute  pressure  after  a  time  /, 

and  to  avoid  complex  mathematical  terms  assume  that  in  the 
expansion  in  the  casing,  in  general  terms, 

Pv  =  K, 

where  v  is  the  volume  of  a  pound  of  steam  and  K  is  a  constant. 
Then  since  v  and  W  are  reciprocals, 

W  =  P  X  constant; 
then  also 

W  =  P,  X  C, 

where  C  is  another  constant. 
In  a  time  dt  we  have  thus 

dW  =  CPdt, 
also 


K 

v 
K. 


GOVERNING  STEAM  TURBINES 


237 


therefore 


, 


KC 


ji/!  =  X"  and 


If  we  take  for  convenience  1/F'  =  2  J  pounds  of  steam  per  second, 
JFt  =  f  pound,  and  Pl  =  165  pounds  per  square  inch  absolute 
pressure,  the  time  required  for  the  average  pressure  in  the  casing 


to  fall  to  100  pounds  is 


3        .        165 
4  X  2.25  100 


159    seconds.     The 


time  required  for  steam  at  165  pounds  absolute  to  fall  to  various 
other  pressures  is  shown  in  Fig.  174. 


200 


.5 

Time-Seconds 


FIG.  174.     Time  Required  for  Pressure  Variations  in  the  Casing  of  a 
Parsons  Turbine. 

With  165  pounds  per  square  inch  absolute  initial  pressure, 
usually  the  no  load  pressure  varies  from  25  to  50  pounds  absolute, 


238 


THE   STEAM  TURBINE 


when,  according  to  the  curve,  the  time  required  to  reach  this 
pressure  without  throttling  is  about  .4  to  .8  second;  and  as  the 
load  is  increased  correspondingly  shorter  times. 

Wilkinson  Governing  Device.  An  important  type  of  valve 
gear  has  been  invented  by  Mr.  James  Wilkinson  and  is  being 
applied  to  the  Wilkinson  turbines.  The  general  arrangement  of 
his  governing  device  is  illustrated  in  Fig.  175  showing  governor, 


FIG.   175.    Wilkinson  Valve  Gear. 

eccentrics,  and  a  series  of  valve  casings.  One  of  these  casings 
contains  what  is  called  a  governor  nozzle  which  is  connected 
mechanically  with  the  eccentrics.  This. is  a  form  of  auxiliary 
valve  of  which  the  function  is  not  primarily  to  discharge  steam 
into  the  turbine  blades  but  to  admit  steam  into  or  eject  it  from 
the  other  valve  casings  for  the  purpose  of  opening  or  closing  them. 
This  governor  nozzle  is  illustrated  in  Fig.  176.  The  important 
feature  of  the  governor  nozzle  is  a  cone-shaped  piston  at  the 
lower  end  of  the  valve  rod  passing  through  the  stuffing-box  and 
connected  to  the  eccentrics  as  shown  in  Fig.  175.  A  cone-shaped 
jet  flows  continuously  over  this  cone.  The  central  chamber  of 
the  governor  nozzle  as  well  as  the  spaces  around  it  except  a  narrow 
annular  passage  communicating  with  a  similar  passage  shown 
by  a  circular  section  in  Fig.  176  at  the  left-hand  side  of  the 


GOVERNING  STEAM  TURBINES 


239 


central  chamber  contains,  in  normal  operation,  steam  at  the 
"admission"  or  initial  pressure. 

The  valves  in  the  other  casings  are  operated  by  the  force 
produced  in  this  annular  chamber  by  the  injector  or  ejector  action 
of  the  cone-shaped  jet.  Steam  to  be  admitted  to  the  annular 
passage  must  pass  around  the  cone-shaped  piston,  and  the 
position  of  this  piston  with  respect  to  the  annular  passage  deter- 
mines the  effective  pressure  of  the  steam  operating  the  admission 
valves.  When  the  cone-shaped  piston  is  in  its  lowest  position 
the  steam  in  passing  around  it  to  enter  the  turbine  nozzles  opposite, 
produces  an  ejector  effect  in  the  annular  passage;  but  when  the 
cone-shaped  piston  is  at  the  other  end  of  its  stroke  the  steam 
produces  an  injector  effect  in  the  annular  passage.  When  the 
injector  effect  predominates  the  pressure  in  the  annular  passage 
is  greater  than  that  in  the  steam  chest,  while  with  the  ejector 
effect  predominating  this  pressure  is  considerably  less. 


FIG.  176.     Governor  Nozzle. 


FIG.  177.     Admission  Valve. 


One  of  the  admission  valves  is  shown  in  Fig.  177.  A  small 
passage  of  circular  section  shown  here  in  the  wall  of  the  steam 
chest  communicates  with  the  governing  valve  or  governor  nozzle. 
This  passage  communicates  with  one  side  of  the  piston  valve 
illustrated  here.  A  spring  is  provided  to  keep  the  valve 
closed  when  the  pressure  in  the  passages  communicating  with 


240  THE   STEAM  TURBINE 

the  governor  piston  is  the  same  as  in  the  steam  chest.  When 
the  pressure  in  the  passages  is,  however,  less,  corresponding  to  the 
ejector  effect,  by  an  amount  greater  than  the  tension  in  the 
spring  (about  25  pounds)  the  valve  is  opened.  With  the  injector 
effect,  on  the  other  hand,  the  valve  will  be  closed. 

All  the  admission  valves  operate  together,  as  the  pressures  are 
approximately  the  same  in  each  of  them.  The  governor  valve 
oscillates  150  times  per  minute.  The  position  of  its  cone-shaped 
valve  with  respect  to  the  annular  passage  communicating  with 
the  admission  valves  is  determined  by  the  height  of  the  governor 
weights  and  is  adjusted  by  means  of  the  levers  shown  in  Fig. 
175.  By  this  arrangement  the  duration  of  the  ejector  effect 
opening  the  valves  is  controlled  by  the  speed. 

All  the  valves  open  a  fixed  number  of  times  in  a  minute,  but  the 
duration  of  the  period  they  are  open  varies  with  the  load. 

Because  the  governor  nozzle  is  always  open  to  the  steam  chest, 
steam  is  never  cut  off  from  this  nozzle,  but  w\ih  careful  designing 
in  proportioning  the  sizes  of  the  nozzles  this  is  no  particular  dis- 
advantage. 

In  its  action  this  valve  gear  is  not  unlike  the  usual  Parsons 
governing  device  which  has  already  been  explained.  It  is  likely 
that  both  are  affected  by  quasi  throttling  at  very  light  loads.  To 
some  extent  the  magnitude  of  this  effect  would  probably  be  in 
proportion  to  the  length  of  the  casing. 

By-pass  Governors.  In  all  turbines  the  area  of  the  steam 
passages  increases  in  going  from  the  high-pressure  end  to  the 
exhaust.  Consequently  it  is  possible  to  pass  a  larger  quantity 
of  steam  through  a  turbine  for  an  overload,  by  admitting  high- 
pressure  steam  into  the  middle  stages  in  addition  to  the  steam 
coming  through  the  high-pressure  nozzles.  This  is  accomplished 
usually  by  the  use  of  an  auxiliary  valve  which  opens  slowly  when 
an  overload  comes  on  the  turbine  and  admits  high-pressure  steam 
directly  into  the  low-pressure  stages.  As  the  steam  entering 
through  the  by-pass  valve  acts  on  fewer  rows  of  blades  than  the 
steam  admitted  under  normal  conditions,  obviously,  of  course,  the 
method  is  uneconomical  and  should,  therefore,  be  used  only  for 


GOVERNING   STEAM  TURBINES 


241 


emergency  loads.  When  a  by-pass  valve  is  used,  the  turbine  is 
designed  to  be  large  enough  to  carry  a  little  more  than  the  normal 
full  load,  at  which,  of  course, 'it  is  most  economical;  and  for 
overloads  it  is  expected  that  the  efficiency  will  be  considerably 
reduced. 

All  the  makers  of  Parsons  and  Rateau  turbines  use  by-pass 
valves.  The  Westinghouse  turbine  has  by-pass  overload  valves 
under  the  control  of  the  governor,  so  that  they  open  automatically 
when  an  overload  comes  on,  but  on  most  turbines  by-pass  valves 
are  opened  by  hand. 

Overload  economy  is  not  usually  of  great  importance,  so  that, 
practically,  it  is  considered  more  feasible  to  use  overload  valves 
than  to  install  additional  turbines.  In  turbines  of  the  Curtis 
type,  which  can  be  made  to  take  a  large  overload  with  the  addition 
of  only  a  few  extra  nozzles  without  increasing  the  other  dimensions, 
by-pass  valves  for  overload  have  no  advantages.  In  Parsons 
turbines,  as  anticipated  in  the  design,  there  is  usually  a  falling 
off  in  speed  when  the  overload  valves  open. 


FIG.  178.     By-pass  Valve  Designed  by  Brown-Boveri  &  Co. 

A  by-pass  governor  is  shown  in  Fig.  178,  which  is  a  diagram- 
matic sketch  showing  the  method  of  admitting  high-pressure 
steam  to  the  low-pressure  stages  of  a  Parsons  turbine.  This 
particular  device  is  due  to  Brown-Boveri  &  Co.  This  design 
shows  the  by-pass  method  applied  to  an  exceptionally  well-made 


242 


THE   STEAM  TURBINE 


turbine  much  used  in  Europe.  The  by-pass  ports  open  only  at 
overload,  and  the  speed  is  regulated  for  small  fluctuations  by 
the  throttling  method.  In  the  figure  the  centrifugal  governor  is 
marked  9,  and  operates  by  means  of  levers  a  balanced  throttle 
valve.  The  by-pass  valve  7,  on  the  other  hand,  is  operated  by 
the  pressure  on  the  piston  10.  Since  this  piston  and  the 
by-pass  valve  are  on  the  same  valve  stem,  they  are  raised  or 
lowered  together  according  as  the  pressure  in  the  steam  chest  is 
high  or  low.  With  a  high  pressure  the  piston  rises,  lifting  with 
it  the  valve  7,  thus  uncovering  the  ports,  shown  on  one  side  of 
the  turbine  at  6,  which  admit  steam  through  the  pipes  3,  4,  and  5 
to  different  parts  of  the  turbine  casing.  Obviously  there  is  a 
considerable  change  in  the  power  developed  immediately  after 
the  steam  is  admitted  to  one  of  the  pipes  3,  4,  or  5;  and  the 
consequent  fluctuation  in  speed  is  taken  care  of  by  the  throttling 
governor  9,  or  by  an  electrical  solenoid  governor  indicated  at  12. 


FIG.  17*9.     By-pass  Valve  Arrangement  for  a  Parsons  Turbine. 

A  simpler  type  of  by-pass  valve  and  governor  arrangement  is 
illustrated  in  Fig.  179.  The  by-pass  valve  is  here  directly  under 
the  control  of  the  governor.  The  governor  is  marked  9  in  the 
figure  and  operates  a  by-pass  piston  valve  7.  The  steam  enters 
the  turbine  through  the  steam  chest  over  the  by-pass  valve. 
When  there  is  no  overload  on  the  turbine,  the  steam  passes 
through  the  side  port,  which  is  shown  open  in  the  figure,  to  the 


GOVERNING   STEAM  TURBINES  243 

steam  space  2  below,  and  from  here  it  passes  through  the  turbine. 
When,  however,  there  is  an  overload,  the  by-pass  valve  7  is  raised 
by  the  governor  and  high-pressure  steam  is  admitted  through 
the  pipe  3  to  some  lower  pressure  stages.  With  still  more  over- 
load the  ports  for  the  pipes  4  and  5  are  also  opened  and  high- 
pressure  steam  is  .admitted  to  stages  intended  for  still  lower 
pressures. 

When  in  order  to  make  repairs  to  the  condenser  equipment, 
or  for  other  reasons,  it  is  necessary  to  run  a  Parsons  turbine 
non-condensing  the  by-pass  valve  is  opened  and  high-pressure 
steam  is  admitted  to  the  intermediate  stage  of  the  turbine.  By 
this  method,  because  of  the  larger  area  of  the  passages,  more 
steam  can  be  used  and  the  turbine  is  able  to  carry  full  load 
without  a  vacuum,  although,  of  course,  at  a  sacrifice  of  economy. 

If  turbines  are  designed  to  take  a  large  overload  without  a 
by-pass,  the  turbine  must  be  of  correspondingly  greater  capacity 
than  the  full  rating  indicates.  The  best  economy  of  steam  will 
then  be  at  the  highest  output,  and  not  quite  so  good  at  three- 
quarter  load  and  full  load.  This  is  the  usual  practice  in  designing 
impulse  turbines;  but  in  the  large  sizes  of  Curtis  turbines  special 
overload  valves  are  provided. 

Curtis  Overload  Valves.  In  Curtis  turbines  of  four  or  five 
stages,  especially  in  the  larger  sizes,  automatic  valves  are  pro- 
vided to  open  additional  nozzles  in  the  diaphragm  between  the 
first  and  second  stages  at  times  of  overload.  The  usual  designs 
of  such  valves  are  similar  to  the  one  shown  in  Fig.  180,  which  is 
arranged  to  operate  when  the  pressure  in  the  first  stage  —  due 
to  a  large  flow  of  steam  —  is  larger  than  the  normal.  This 
design  consists  essentially  of  a  piston  valve  fitted  with  a  spring  of 
sufficient  strength  to  balance  the  unequal  pressures  on  its  faces 
for  normal  operation  of  the  turbine.  In  the  position  shown  in 
the  figure  the  valve  is  closed;  that  is,  no  steam  passes  through 
it  from  the  first  stage  to  the  nozzles  discharging  into  the  second 
stage.  The  pressure  on  its  upper  face  is  that  in  the  first  stage, 
while  the  pressure  on  the  lower  face  is  approximately  that  in 
the  third  stage. 


244 


THE  STEAM  TURBINE 


Stage 


Second  6&3ffe 


Th/rcf Stage 


FIG.  180.     Curtis  Overload  "Stage"  Valve. 


GOVERNING  STEAM  TURBINES 


245 


As  the  flow  of  steam  increases  due  to  increasing  the  load  with 
a  constant  nozzle  area  the  pressure  will  become  greater  in  each 
of  the  stages;  but  obviously  the  pressure  will  be  increased  much 
more  in  the  first  stage  than  in  the  other  stages,  and  at  an  overload 
the  difference  in  pressure  between  the  first  and  third  stages 
becomes  great  enough  to  overcome  the  resistance  of  the  spring 
on  the  overload  or  stage  valve,  so  that  it  is  forced  from  its  seat. 
Steam  then  passes  through  a  port  communicating  with  an  extra 
set  of  nozzles  discharging  into  the  second  stage. 

These  valves  should  be  adjusted  by  varying  the  tension  of  the 
spring,  so  that  they  tend  to  open  and  close  within  a  compara- 
tively small  range  of  first-stage  pressure. 

If  the  adjustment  of  such  a  valve  has  not  been  properly  made 
and  the  valve  remains  open  with  a  load  fluctuating  in  a  wide 
range  between  overload  and  considerably  less  than  normal,  the 
economy  may  be  seriously  impaired,  or  if  one  of  these  valves 
remains  in  a  partly  closed  position  so  as  to  throttle  the  steam,  the 
economy  will  be  affected  at  all  loads.  Such  valves  to  be  efficient 
should  open  and  close  abruptly. 

Experimental  Data  Concerning  Governing.  Something  should 
be  said  about  the  experimental  results  at  hand  concerning  the 
different  methods  of  governing.  Curves  illustrating  the  effects 
of  throttling  have  been  shown  in  Fig.  161,  but  a  more  satisfac- 
tory comparison  can  be  made  from  the  following  table.* 


Rated  Full 
Load. 

Fraction  of  Load. 

Kw. 

i 

1 

Full. 

Curtis  

$°° 
600 

20 
2OO 
500 
500 
400 
1250 
350 

5*-7 

52-4 
68.0 

52-3 
56.0 

S5-o 
57-o 

57-3 
58.8 

76.2 

76.5 
82.0 
76.2 
78.0 
77.2 
78.5 
78.5 
80.3 

TOO 
100 
IOO 
IOO 
IOO 
IOO 
IOO 
IOO 
IOO 

Curtis  

De  Laval.  

De  Laval  

C.  A.  Parsons  Company  

Rateau  

Westinghouse-Parsons  

Westinghouse-Parsons  

Zoelly  

*  Mechanical  Engineer,  Jan.  20,  1906. 


246  THE  STEAM  TURBINE 

Fractions  of  load  given  at  the  top  of  each  column  refer  to 
fractions  of  the  most  efficient  load.  Steam  consumption  at  the 
different  loads  is  expressed  as  a  percentage  compared  with  the 
steam  consumption  at  the  most  economical  load  for  each  partic- 
ular machine.  In  other  words,  if  the  economy  of  any  of  these 
turbines  were  as  good  'at  half  load  as  at  full  load  we  should  have 
in  the  table  under  the  column  for  one-half  load  50  per  cent,  etc. 

Results  in  this  table  must  be  used  guardedly  and  not  confused 
•with  steam  consumption.  For  example,  the  De  Laval  20okilo- 
watt  turbine  appears  to  such  good  advantage  here  because  the 
system  of  governing  used  for  these  tests  was  nearly  ideal.  Full 
load  steam  consumption,  on  the  other  hand,  was  high  com- 
pared with  any  other  make  of  turbine  in  the  list.  The  Zoelly, 
Rateau,  and  the  2O-kilowatt  De  Laval  turbines  used  simple 
throttling  governors. 

The  Curtis  and  the  2oo-kilowatt  De  Laval  were  governed  by 
varying  the  number  of  nozzles  to  suit  the  load.  The  original 
Parsons  and  Westinghouse-Parsons  turbines  used  the  " blast'' 
governor.  These  data  lead  to  the  conclusion  that  the  Curtis 
and  the  experimental  De  Laval  (200  kilo  watt)  give  the  best 
results  as  regards  the  method  of  governing. 

Sufficient  data  are  not  available  of  the  performance  of  Wil- 
kinson and  Allis-Chalmers  turbines  to  be  included  in  this  com- 
parison. 


CHAPTER  IX. 
LOW-PRESSURE   STEAM  TURBINES. 

EARLY  in  the  period  of  steam  turbine  development  it  became 
apparent  that  these  new  types  of  prime  movers  were  capable  of 
operating  with  ratios  of  expansion  far  beyond  those  economically 
possible  with  reciprocating  engines. 

In  the  discussion  of  the  effect  of  vacuum  on  steam  consump- 
tion the  good  results  obtained  with  turbines  running  at  a  high 
vacuum  were  clearly  shown.  With  a  high  vacuum  the  heat 
efficiency  of  a  reciprocating  engine  is  not  nearly  so  good  as  that 
of  a  turbine,  because  it  is  not  desirable  to  make  the  engine  cyl- 
inders large  enough  to  handle  economically  the  great  volume  of 
steam  we  have  to  deal  with  when  the  exhaust  pressure  is  very 
low.  For  pressures  slightly  above  atmospheric,  however,  a  first- 
class,  slow-speed  reciprocating  engine  has  a  slight  advantage 
over  the  turbine.  We  can  see,  then,  that  a  combination  of  a 
non-condensing  reciprocating  engine  with  a  condensing  turbine, 
the  latter  taking  exhaust  from  the  former,  might  well  be  sug- 
gested. 

Fig.  181  shows  graphically  the  volumes  of  the  steam  in  each 
of  the  five  stages  of  a  Curtis  turbine  and  illustrates  how  rapidly 
the  volume  increases  at  very  low  pressures.  Reciprocating 
engines  may  be  designed  to  operate  with  improved  economy  up 
to  25  or  26  inches  vacuum;  but  this  is  about  the  limit.  Steam 
turbines,  on  the  other  hand,  will  operate  economically*  with 
steam  at  the  highest  vacuum  practically  obtainable. 

The  initial  pressure  of  low-pressure  steam  turbines  is  usually 
that  of  the  exhaust  steam  from  non-condensing  engines.  With 

*  It  is  not  always  commercially  profitable  to  design  a  plant  for  operation  at  an 
extremely  high  vacuum,  as  the  first  cost  of  condensers  and  auxiliaries  is  usually  a 
deciding  factor. 

247 


248 


THE    STEAM   TURBINE 


steam  admitted  to  the  engine  at  200  pounds  and  exhausted  from 
the  turbine  at  28  inches  vacuum,  theoretically  there  is  no  differ- 
ence in  the  total  economy  of  a  unit  consisting  of  a  reciprocating 


11 

U    C 


«> 

Is. 


-3 


3  .s 


s  a 


^ 


£  - 

'' 


engine  operating  with  an  exhaust  turbine  taking  steam  within 
the  Umits  from  7  pounds  to  15  pounds  per  square  inch  absolute.* 
The  curve  in  Fig.  i8ia  shows  the  energy  (B.T.U.)  made  available 

*  Proc.  Inst.  of  Naval  Architects,  1908. 


LOW-PRESSURE    STEAM   TURBINES 


249 


by  expansion  from  26  inches  vacuum  to  higher.     The  rapid  in- 
crease above  27  inches  is  an  important  consideration  and  makes 


29 


28 


27 


28 


0  10  20  30  40  50  60  70 

Energy  in  B.T.U.  available  between  26"  and  29"  vacuum. 

FIG.  i8ia.     Curve  Showing  Enormous  Energy  Made  Available  for  Work  by 
Expansions  to  High  Vacuums. 

it  very  essential  to  get  a  very  high  vacuum  in  the  exhaust  on 
account  of  the  enormous  energy  obtainable. 

Flattening  out  of  the  curve  of  the  specific  volumes  of  steam, 
Fig.  i8ib,  is  also  interesting  in  this  connection.  The  increase  in 
volume  as  the  expansion  is  carried  beyond  28  inches  of  vacuum 
is  not  generally  realized. 

The  following  table  prepared  by  the  General  Electric  Company 
shows  the  large  amount  of  work  that  is  made  available  by  the 
use  of  turbines  in  connection  with  existing  non-condensing  and 
condensing  plants,  the  steam  being  delivered  to  the  steam  en- 
gine at  165  pounds  per  square  inch  absolute  pressure. 

Owing  to  the  rapid  development  of  the  turbine  industry  for 
high  speed  work  and  the  close  attention  to  this  branch  of  turbine 
applications  required  of  designers,  the  "  combination  "  system  of 
reciprocating  engines  and  turbines  was  comparatively  neglected. 
Only  recently  the  advantages  of  this  system  have  come  to  be 


25° 


THE   STEAM   TURBINE 


generally  recognized,  and  particularly  in  connection  with  marine 
propulsion.  Parsons  has  never  advised  the  installation  of  an 
"  all  turbine  "  arrangement  for  ships  designed  for  a  speed  of 


Atmos- 

pheric 

Inches  of  Vacuum. 

Pressure. 

Pressure  of  steam  at  tur-j 

bine  admission  valve  [ 

O 

4" 

8 

12 

16 

20 

24 

in  inches  of  vacuum.    J 

Per   cent,    gained    over 

output  of  engine  when 
worked  with  high  vac- 
uum, the  turbine  ex- 

26.1 

26.5 

26.8 

26.3 

25-3 

23-6 

20 

hausting  to  a  vacuum 

of  28^  inches. 

less  than  1 5  knots,  and  for  moderate  or  slow  speeds  his  designers 
have  recommended  the  "  combination  "  system.  According  to 
one  of  his  designs  for  a  cargo  vessel  intended  for  a  speed  of  nj 


29 


28 


27 


26 


100 


200 


500 


300  400 

Cubic  Feet  per  Pound. 
FIG.  i8ib.     Curve  of  Specific  Volumes  between  26  and  29  Inches  of  Vacuum. 

knots,  if  provided  with  a  reciprocating  engine  discharging  steam 
into  the  turbine  at  7  pounds  per  square  inch  absolute  pressure, 
the  steam  consumption  was  estimated  to  be  15  to  20  per  cent,  less 


LOW-PRESSURE    STEAM   TURBINES 

than  that  of  an  "all  turbine"  arrangement,  or  of  triple  expansion 
engines  of  the  type  usually  fitted  to  this  class  of  vessel.  The 
"  combination  "  system  gives  a  vessel  also  greater  maneuvering 
power  than  if  driven  only  by  turbines. 

On  land  low-pressure  turbines  have  been  installed  principally 
in  connection  with  rolling  mill  engines  in  steel  works  and  wind- 
ing engines  in  mines.  In  both  cases  the  engines  are  stopped 
or  are  running  practically  idle  a  large  part  of  the  time.  These 
engines  are  usually  reversing  and  are  operated  non-condensing. 
When  a  low-pressure  steam  turbine  is  installed  to  take  the 
exhaust  from  such  engines  an  equal  amount  of  power  can  be 
obtained  from  the  turbine  (at  28  inches  vacuum)  as  from  the 
engine,  thereby  doubling  the  power  of  the  plant  without  increas- 
ing the  consumption  of  coal  or  the  size  of  the  boiler  plant. 

Throughout  the  country  there  are  a  great  many  reciprocating 
engines  exhausting  into  the  atmosphere,  and  the  exhaust  from 
these  engines  is  often  wasted.  There  is  no  doubt  that  when  these 
plants  need  increased  capacity  an  installation  of  exhaust  turbines 
will  be  profitable,  even  in  most  cases  where  there  is  no  supply 
of  water  for  condensing,  and  cooling  towers  must  be  erected. 

There  are  also  many  power  plants  equipped  with  high  grade 
compound  reciprocating  engines  operating  condensing  which  have 
a  highei  efficiency  (not  steam  consumption)  when  operating  non- 
condensing,  or  at  a  comparatively  small  vacuum,  than  when 
operating  condensing.  In  such  cases  the  installation  of  low- 
pressure  turbines  is  probably  always  profitable.  As  an  instance 
of  the  uses  of  exhaust  steam  turbines  the  following  paragraph 
is  quoted  from  a  report  prepared  by  a  company  manufactur- 
ing large  sizes  of  both  reciprocating  steam  engines  and  steam 
turbines. 

•  "  A  compound  reciprocating  engine  with  cylinder  ratios  of 
3.5  :  i,  say  of  diameters  28  inches  and  52  inches,  with  150  pounds 
initial  pressure,  may  be  assumed  to  have  1000  kilowatts  econom- 
ical capacity  when  running  condensing  and  having  a  steam  con- 
sumption of  about  22  pounds  per  kilowatt-hour.  This  engine 
if  operated  non-condensing  should  have  valve  gears  adjusted  to 


25ob  THE   STEAM  TURBINE 

develop  1700  I.H.P.,  when  it  would  consume  about  20  pounds  of 
steam  per  I.H.P.  per  hour.  This  gives  30,600  pounds  steam 
available  for  the  turbine,  allowing  10  per  cent,  of  moisture  in  the 
exhaust  of  the  reciprocating  engine.  The  total  amount  of  steam 
passing  the  reciprocating  engine,  however,  being  34,000  pounds, 
30,600  pounds  would  develop  not  less  than  1073  brake  horsepower 
in  the  turbine.  Allowing  94  per  cent,  for  the  mechanical  effi- 
ciency of  the  reciprocating  engine,  the  combined  horsepower 
developed  would  be  2673  brake  horsepower  and  the  steam  con- 
sumption of  the  two  units  12.7  pounds  per  brake  horsepower,  or 
1 8  pounds  per  kilowatt-hour,  which  is  a  remarkable  performance 
for  engines  of  such  capacities  operating  without  superheat. 
Compared  with  the  performance  of  the  reciprocating  engine 
running  condensing,  this  gives  75  per  cent,  increase  of  power 
and  1 8  per  cent,  saving  of  steam." 

Fig.  i8ic  shows  a  very  important  installation  of  5Ooo-kilo- 
watt  steam  turbine-generators  in  combination  with  reciprocating 
engines  of  the  same  power  rating.  A  low-pressure  steam  tur- 
bine has  been  installed  to  take  the  exhaust  from  each  engine. 

Condensing  engines  when  changed  to  non-condensing  operation 
do  not  necessarily  have  their  capacity  in  horsepower  reduced 
because  of  the  great  increase  of  back  pressure  against  which 
they  must  then  operate.  Such  reduction  would  appear,  however, 
on  first  thought  to  be  the  natural  result;  but,  contrarily,  the 
capacity  of  such  an  engine  when  changed  to  non-condensing 
operation  may  be  unaltered  or  even  in  exceptional  cases  may  be 
actually  increased;  particularly  is  this  the  case  if  the  engine  is 
one  designed  for  a  high  expansion  ratio.  Under  these  conditions 
the  high-pressure  cylinder  must  have  enough  volume  to  pass 
the  required  amount  of  steam,  without  having  the  cut-off  come 
so  late  as  to  sacrifice  all  opportunity  to  use  the  steam  with  a 
reasonably  good  expansion.  There  is  an  interesting  reason  for 
the  capacity  of  many  compound  engines  not  being  reduced  when 
this  change  is  made.  In  this  adjustment  the  cut-off  of  the  high- 
pressure  cylinder  has  been  shifted  to  make  it  late  enough  so  that 
expansion  in  the  low-pressure  cylinder  will  not  cause  a  loop  in 


LOW-PRESSURE   STEAM   TURBINES 


2500 


FIG.  i8ic.     Steam  Turbines  taking  the  Exhaust  from  Large  Reciprocating  Engines. 


25od  THE   STEAM  TURBINE 

the  indicator  diagram  or  to  a  final  pressure  in  the  low-pressure 
cylinder  when  its  exhaust  valve  opens,  which  is  lower  than  the 
average  pressure  in  the  exhaust  line  supplying  the  turbine, 
which  is  also  the  engine  exhaust  pipe.  As  the  result  of  this 
adjustment  of  cut-off  in  the  high-pressure  cylinder  the  top  part 
of  the  indicator  cards  taken  from  it  will  be  observed  to  be  much 
enlarged  and  of  greater  area  than  before,  the  increase  being  in 
some  cases  greater  even  than  the  area  which  is  lost  at  the  bottom 
of  the  low-pressure  diagram  by  raising  its  exhaust  pressure  to 
about  atmospheric.  But  there  are  also  many  compound  engines 
operating  condensing  in  which  the  release  in  the  low-pressure 
cylinder  occurs  when  the  pressure  is  relatively  high,  possibly  as 
high  as  atmospheric.  Now  the  application  of  a  low-pressure 
turbine  to  take  the  exhaust  from  this  engine  would  have  the 
effect  of  very  materially  reducing  the  capacity  of  the  engine,  as 
the  benefits  to  be  obtained  to  avoid  the  loop  in  the  low-pressure 
diagram  have  been  sacrificed  in  the  original  design  of  the  engine 
and  there  is  no  chance  to  increase  the  area  of  the  top  of  the 
diagram. 

Engines  designed  to  operate  both  condensing  and  non-con- 
densing have  generally  the  valves  adjusted  so  that  normally 
there  will  be  a  rather  high  back  pressure  at  the  point  of  release 
in  the  low-pressure  cylinder,  so  that  this  " looping"  of  its  indi- 
cator diagram  will  be  avoided  when  running  non-condensing  or 
on  light  loads.  When  carrying  full  loads  or  an  overload,  on 
the  other  hand,  the  expansion  will  not  be  complete  and  a  serious 
loss  results  in  such  engines  designed  for  operation  with  loads 
varying  considerably.  In  this  case  when  the  low-pressure  tur- 
bine is  applied  there  will  be  no  occasion  for  the  high  back  pressure 
at  release,  and  the  valve  setting  can  be  changed  to  exhaust  the 
steam  from  the  low-pressure  cylinder  when  running  at  full  load 
so  that  this  pressure  will  be  about  atmospheric  or  possibly  a 
half  pound  or  a  pound  above  to  assist  in  getting  the  steam  readily 
through  the  exhaust  ports  of  the  engine.  For  this  condition  the 
blades  of  the  low-pressure  turbine  taking  this  exhaust  should  be 
designed  for  initial  pressure  approximately  atmospheric  when 


LOW-PRESSURE   STEAM  TURBINES  2506 

getting  the  amount  of  steam  used  by  the  engine  at  full  load. 
Now  when  the  cut-off  is  shifted  back  by  the  governor  to  keep  the 
inlet  valve  of  the  engine  open  only  half  as  long  as  at  full  load, 
only  half  as  much  steam  will  be  delivered  to  the  turbine  and  the 
absolute  inlet  pressure  to  it  will  also  be  reduced  to  half  its 
former  value.  The  final  result  is  that  the  expansion  of  the  steam 
in  the  low-pressure  cylinder  of  the  engine  will  be  about  to  the 
inlet  pressure  of  the  turbine  and  there  will  be  no  appreciable 
"  looping  "  of  its  indicator  diagram,  indicating  that  the  condi- 
tions as  regards  effective  expansion  of  the  steam  are  excellent. 

In  large  power  plant  installations,  it  is  the  best  practice  to 
adhere  to  the  "  unit  system  "  throughout;  that  is,  in  providing 
a  separate  low-pressure  turbine  for  each  engine  and  also  a  sep- 
arate condenser  for  each  turbine.  This  method,  although  con- 
siderably more  expensive  than  that  of  passing  the  exhaust  steam 
from  several  engines  to  a  single  receiver  supplying  a  relatively 
larger  low-pressure  turbine  is,  however,  much  to  be  desired  as 
it  gives  so  much  greater  flexibility  in  the  operation  of  the  plant 
and  reduces  very  much  the  liability  to  an  enforced  shut-down 
of  the  plant  due  to  condenser  troubles.  Maximum  load  on  a 
low-pressure  turbine  connected  to  the  engine  as  here  described, 
that  is,  without  having  a  governor  on  its  steam  supply  pipe,  is 
usually  reached  in  conventional  designs  when  the  pressure  at 
the  inlet  to  the  turbine  is  about  twenty  pounds  per  square  inch 
absolute.  A  relief  valve  must  always  be  provided  hi  this  low- 
pressure  supply  line  which  should  open  to  let  steam  out  to 
atmospheric  exhaust  when  this  pressure  is  exceeded.  Steam 
discharged  through  the  relief  valve  is  an  excess  above  what 
can  be  used  in  the  turbine  and  is  obviously  wasted. 

All  possible  precautions  should  be  taken  to  prevent  the  leak- 
age of  air  into  all  that  part  of  the  system  operating  at  less  than 
atmospheric  pressure.  This  air  is  very  detrimental  to  the  proper 
action  of  a  good  condenser  as  it  reduces  the  attainable  vacuum 
and  consequently  renders  impossible  the  full  gains  to  be  ex- 
pected from  the  installation  of  low-pressure  turbines.  Such 
leaks  occur  most  generally  in  the  joints  of  the  exhaust  piping  of 


25of  THE   STEAM   TURBINE 

both  engine  and  turbine,  also  in  imperfectly  tight  relief  valves, 
as  well  as  through  the  stuffing-boxes  on  the  piston  rods  of  the 
low-pressure  cylinder  of  compound  engines.  To  eliminate  these 
difficulties  the  piping  should  be  examined  and  tested  frequently 
by  applying  a  lighted  taper  to  all  questionable  joints  to  observe 
whether  the  vacuum  inside  the  piping  tends  to  draw  the  flame 
toward  it,  as  would  occur  if  there  were  a  leak.  Another  pre- 
caution often  necessary  is  to  put  a  special  type  of  stuffing  box  on 
the  piston  rods  of  the  engines,  these  boxes  being  supplied  with 
steam  at  a  pressure  slightly  above  atmospheric  so  that  air  leakage 
inward  is  prevented. 

When  several  engines  are  connected  up  to  supply  exhaust  to 
a  single  low-pressure  turbine,  it  is  always  most  desirable  to  turn 
over  each  engine  for  several  revolutions  with  the  piping  connec- 
tions arranged  so  that  the  engine  just  starting  will  exhaust  into 
the"  atmosphere.  This  should  be  done  in  order  to  avoid  dis- 
charging the  air  in  the  engine  cylinders  into  the  low-pressure 
turbine  and  into  the  condenser.  If  this  precaution  is  not  taken 
when  several  engines  are  operating  and  another  is  started  the 
effect  in  vacuum  reduction  will  be  observed. 

In  matters  of  design,  the  low-pressure  steam  turbine  presents 
no  new  problems.  In  fact  its  construction  is  in  many  respects 
simpler  because  of  requiring  fewer  complicated  details  than  the 
usual  types  of  high-pressure  or  "  complete  expansion  "  turbines 
that  have  been  studied.  The  relatively  short  length  of  the 
shaft  or  drum  required  for  low-pressure  turbines  makes  for 
rigidity  and  freedom  from  vibration  stresses.  The  skill  of  the 
engineer  comes  into  play  in  this  new  field  almost  entirely  in  the 
methods  of  application  to  conditions  that  a  few  years  ago  were 
not  thought  of.  The  primary  consideration  in  practically  all 
these  applications  is  to  utilize  as  much  as  possible  of  the  avail- 
able exhaust  steam  about  the  plant,  either  in  the  low-pressure 
turbine  or  in  some  still  more  efficient  method.  While  it  is  the 
object  to  show  here  the  great  advantages  of  this  type  of  prune 
mover,  yet  it  should  be  pointed  out  that  there  are  often  condi- 
tions arising  in  power  plant  practice  when  exhaust  steam  can 


LOW-PRESSURE   STEAM  TURBINES  25Og 

be  used  much  more  efficiently  than  in  any  known  type  of  prime 
mover.  The  maximum  thermal  efficiency  of  a  low-pressure  tur- 
bine cannot  well  be  made  to  greatly  exceed  10  per  cent.,  and  even 
with  this  low  efficiency  when  exhaust  steam  is  a  by-product,  with 
no  other  available  use,  the  addition  of  such  a  turbine  to  the  plant 
will  in  such  cases  produce  a  great  saving  in  coal  bills.  In  other 
cases,  however,  where  exhaust  steam  can  be  used  advanta- 
geously for  heating  water,  as  for  example  in  feed-water  heaters, 
hot-water  vats  in  manufacturing  processes,  or  for  heating  build- 
ings, it  would  certainly  be  false  economy  to  use  the  exhaust 
steam  in  a  turbine  and  install  low-pressure  boilers  for  heating 
water.  In  the  cases  cited  the  thermal  efficiency  of  the  process 
of  heating  water  with  low-pressure  steam  is  usually  about  80 
per  cent.,  which  is  to  be  compared  with  the  10  per  cent,  efficiency 
of  the  turbine.  In  this  analysis  it  must  not  be  overlooked,  how- 
ever, that  a  steam  power  plant  operating  non-condensing  can 
use  only  a  very  small  percentage  of  the  total  amount  of  its 
exhaust  steam  for  heating  the  feed  water.  In  fact  in  the  aver- 
age steam  engine  plant  operating  non-condensing  only  about 
one-sixth  of  the  exhaust  steam  can  be  used  for  heating  the  feed 
water,  the  other  five-sixths  being  discharged  into  the  atmos- 
phere through  the  exhaust  head. 

Of  first  importance  among  the  general  considerations  affecting 
low-pressure  turbine  applications  is  the  providing  of  adequate 
facilities  for  the  removal  of  water  and  oil  from  the  steam  before 
it  enters  the  turbine.  Very  wet  steam  can  have  no  considerable 
deleterious  effect  on  the  turbine  compared  with  the  disastrous 
results  often  experienced  in  steam  engine  practice.  It  has  the 
effect,  however,  of  increasing  enormously  the  fluid  friction  in  the 
turbine  blades  and  therefore  of  reducing  the  output  and  raising 
the  steam  consumption.  Oil,  when  clean  and  pure,  is  not  nec- 
essarily objectionable  in  the  turbine  and  will  pass  through  with- 
out accumulating,  but  in  cases  where  boilers  sometimes  foam 
and  discharge  sulphates  and  carbonates  with  the  steam,  these 
will  mix  with  the  oil  and  form  a  gummy  deposit  on  the  blades. 
This  deposit  is  not  ordinarily  removed  by  erosion,  particularly 


THE   STEAM   TURBINE 


when  steam  velocities  are  not  over  400  feet  per  second,  and  will 
often  choke  the  steam  passages  between  the  blades. 

Until  recently  in  America  exhaust  steam  turbines  were  usually 
arranged  to  take  steam  directly  from  the  exhaust  pipe  of  the 
engines  without  intervening  valves  or  governing  mechanisms. 
A  generator  direct  connected  to  the  turbine  will  operate  very 
satisfactorily  with  generators  adapted  for  connection  in  parallel 
to  engine-driven  generators,  and  the  turbine  set  thus  "  floats 
on  the  system."  As  it  receives  only  steam  exhausted  from  the 
engine  its  output  will  therefore  vary  as  the  load  on  the  engine. 
When  the  load  becomes  light  the  steam  supply  will  be  reduced 
by  the  governor  on  the  reciprocating  engine.  In  the  case  of  direct 
current  units  the  generators  may  have  shunt  windings,  and  as  the 
voltage  will  vary  nearly  as  the  speed,  the  load  will  be  automati- 
cally proportioned  between  the  reciprocating  and  turbine  units. 


FIG.  i8id.     Simplest  Combination  of  Low-pressure  Steam  Turbine  and  Recip- 
rocating Steam  Engine. 

The  most  common  and  probably  also  the  simplest  application 
of  the  low-pressure  turbine  is  shown  in  Fig.  i8id.    As  shown  the 


LOW-PRESSURE   STEAM   TURBINES  250! 

generators  connected  to  both  the  reciprocating  engine  and  the 
turbine  are  connected  to  the  same  three-phase  alternating-current 
circuit  and  here  also  no  governor  is  used  on  the  turbine.  By 
this  arrangement  the  turbine  will  take  automatically  its  share 
of  the  total  electrical  load  in  proportion  to  the  amount  of  steam 
supplied  to  it.  If  it  tends  to  forge  ahead  of  the  reciprocating 
unit  it  will  take  more  of  the  load,  leaving  less  for  the  engine 
whose  speed  will  immediately  increase  until  its  governor  re- 
duces the  flow  of  steam  to  both  the  reciprocating  engine  and  to 
the  turbine,  thus  controlling  with  one  governor  the  amount  of 
steam  supplied  to  the  complete  system.  In  case  the  generators 
driven  by  the  engine  and  turbine  are  of  the  direct  current  type, 
as  the  turbine  forges  ahead,  and  takes  more  load,  the  increase 
in  its  speed  raises  the  voltage  slightly,  which  puts  more  current 
through  the  fields  of  the  generators  and  tends  to  reduce  the 
speed.  Self -regulation  is  thus  admirably  accomplished.  Obvi- 
ously for  the  same  reason  it  is  possible  to  vary  the  speed  of  the 
turbine  slightly  by  adjusting  the  field  rheostat. 

The  quantity  of  steam  used  determines  obviously  the  relative 
amount  of  load  carried  by  the  low-pressure  turbine;  the  greater 
the  amount  of  steam  the  greater  the  proportion  of  load  taken 
by  the  turbine,  which  is  due,  of  course,  to  the  variation  of  pres- 
sure in  the  receiver.  As  this  pressure  increases  the  total  range 
of  pressure  available  for  the  turbine  increases,  the  heat  available 
per  pound  of  steam  increases,  and  consequently  more  work  is 
done,  assuming,  of  course,  a  constant  vacuum  in  the  turbine 
exhaust. 

In  the  method  of  low-pressure  turbine  installation  described 
in  the  preceding  paragraphs,  where  the  turbine  operated  with- 
out a  governor  of  its  own,  the  electrical  machines  driven  (gen- 
erators) were  of  similar  types;  that  is,  all  the  current  generated 
was  supplied  to  a  single  line.  It  is  not  infrequent,  however,  for 
low-pressure  steam  turbines  to  be  installed  to  operate  with  re- 
ciprocating engines  in  power  houses  where  the  generators  on  the 
engines  are  to  supply  direct  current  lines  and  the  generators  on 
the  turbines  are  to  supply  alternating  current  for  transmission 


THE   STEAM  TURBINE 


to  a  distance.  Such  an  arrangement  is  shown  in  Fig.  i8je. 
Obviously  the  engine  and  the  turbine  must  each  have  its  own 
governor.  If  the  low-pressure  turbine  were  arranged  to  take 
only  the  load  from  the  alternating  current  line  there  would  be 
much  steam  wasted  when  the  direct  current  load  happened  to  be 
heavy  and  the  other  light;  and  conversely' it  might  be  neces- 
sary to  sometimes  supply  the  low-pressure  turbine  with  high- 


FIG.  i8ie.     Application  of  a  Rotary  Converter  in  the  Combination  of  Low- 
pressure  Turbine  with  Reciprocating  Engine. 

pressure  steam  when  the  load  on  the  engine  was  light.  A  most 
satisfactory  method  to  avoid  these  difficulties  is  to  install  a 
rotary  converter  or  motor-generator  set  as  shown.  Any  in- 
equality of  the  two  loads  will  then  be  taken  care  of  and  the  load 
coming  on  the  engine  and  on  the  turbine  will  be  divided  auto- 
matically to  give  the  best  results.  This  sort  of  arrangement 
might  not  be  very  satisfactory  for  taking  care  of  electric  lighting 
loads  if  there  were  likely  to  be  exceptionally  frequent  reversals 
of  the  operation  of  the  converter  from  alternating  to  direct 
current  and  vice  versa,  as  there  might  be  a  voltage  change  of 
several  per  cent,  which  would  perceptibly  affect  illumination 
until  again  adjusted  at  the  switchboard.  In  most  cases,  how- 


LOW-PRESSURE   STEAM  TURBINES 


ever,  the  demand  for  one  kind  of  current  will  always  predomi- 
nate, so  that  this  sort  of  reversal  is  not  likely  to  be  troublesome. 
Another  application  of  a  low-pressure  turbine  is  illustrated  in 
Fig.  1 8 if,  where  the  steam  engine  drives  a  line  of  shafting  through 
a  belt  drive  and  the  low-pressure  steam  exhausted  from  the 
engine  goes  to  a  turbine  generator  unit  supplying  an  electrical 
transmission  line.  In  this  case  the  engine  and  the  turbine  must 
each  have  a  governor,  as  the  loads  on  the  two  machines  are  en- 


A.C.Line 


FIG.  i8if.     Application  of  Condenser  in  Combination  System. 

tirely  unrelated.  The  device  adopted  in  this  case  to  make  the 
plant  as  economical  of  steam  as  possible  is  to  operate  the  engine 
with  the  excess  of  steam,  above  that  required  for  the  low-pressure 
turbine,  to  be  discharged  directly  into  the  condenser.  By  this 
method  the  engine  will  operate  at  times  at  a  fairly  good  vacuum 
as  determined  by  the  relative  amounts  of  steam  and  absolute 
pressures  in  lines  A  and  B.  The  governor  on  the  turbine  oper- 
ates only  the  by-pass  valve  V,  regulating  the  flow  of  steam  from. 


2 


THE   STEAM   TURBINE 


the  engine  exhaust  into  the  condenser.  When  the  load  on  the 
turbine  is  light  this  valve  will  be  nearly  wide  open,  deflecting 
only  a  small  amount  of  steam  into  the  turbine;  but  when  the 
turbine  load  gets  near  its  full  capacity,  the  condenser  by-pass 
valve  will  be  nearly  closed.  It  becomes  thus  possible  for  the 
engine  to  obtain  the  advantages  of  nearly  full  vacuum  when 
the  turbine  is  running  light.  It  is  a  good  practice  to  put  the 
usual  type  of  valve  on  the  inlet  pipe  to  the  turbine  which  will 
also  be  controlled  by  the  governor  to  prevent  the  turbine  run- 
ning away  on  very  light  load. 

A  very  interesting  type  of  installation  is  shown  in  Fig.  i8ig. 


FIG.  i8ig.     Application  of  Synchronous  Motor  in  Combination  System. 

The  method  illustrated  here  consists  in  the  installation  of  an 
electric  motor  of  the  synchronous  type  supplied  with  current 
from  the  generator  driven  by  the  turbine  and  having  the  pulley 
on  its  shaft  belted  to  the  line  shafting  driven  by  the  engine.  In 
case  the  electrical  load  on  the  turbine-generator  set  becomes  too 
large  for  it  to  handle  it  will  slow  down  slightly  with  the  result 
that  the  synchronous  motor  will  be  driven  from  the  line  shafting 


LOW-PRESSURE   STEAM   TURBINES  251 

as  a  generator  to  supply  more  current  to  the  electrical  supply 
lines.  The  additional  load  coming  on  the  engine  as  the  result 
of  driving  the  motor  will  cause  the  governor  to  open  the  inlet 
valve  wider  on  the  engine  and  admit  a  larger  amount  of  steam 
to  the  system.  Conversely,  when  the  line  shafting  is  overloaded 
the  governor  on  the  engine  admits  more  steam  to  the  system,  in 
greater  amount,  however,  than  the  turbine  requires.  This  re- 
sults in  a  speeding  up  of  the  turbine  and  a  forging  ahead  of  the 
synchronous  motor  so  that  it  acts  now  purely  as  a  motor  to 
assist  in  driving  the  shafting. 

Provisions  for  Intermittent  Supply  of  Steam.  An  ingenious 
development,  largely  due  to  Professor  Rateau,  has  been  applied 
to  cases  where  the  supply  of  exhaust  steam  is  intermittent,  as 
in  the  case  of  rolling  mill  and  winding  engines.  Rateau's  device, 
called  an  accumulator,  is  used  to  bridge  over  the  "dead  periods," 
and  by  providing  sufficient  capacity  it  can  be  made  to  provide  a 
practically  constant  supply  for  an  exhaust  turbine. 

Rateau's  Accumulator.  This  regenerator  or  accumulator  is 
shown  in  Fig.  182,  illustrating  longitudinal  and  transverse  sec- 
tions. This  regenerator  consists  of  a  large  cylindrical  shell  partly 
filled  with  water.  When  the  engine  exhausting  into  it  is  running 
the  steam  is  delivered  as  a  spray  through  the  small  holes  in  a  num- 
ber of  pipes  immersed  in  the  water.  By  this  method  some  of  the 
steam  is  condensed  and  gives  up  heat  to  the  most  of  the  water. 

As  these  accumulators  operate  usually  with  steam  at  atmos- 
pheric pressure,  the  entering  steam  will  have  a  temperature  of  212 
degrees  F.  and  will  tend  to  heat  the  water  to  that  temperature.  If, 
now,  the  engine  stops,  the  supply  of  exhaust  steam  is  discontinued, 
and  the  flow  of  steam  to  the  turbine  will  tend  to  make  the  pres- 
sure fall  off  slightly  so  that  212  degrees  F.  will  then  be  slightly 
above  the  temperature  of  boiling  water  at  this  lower  pressure.  In 
this  way  the  water  will  be  evaporated  to  supply  steam  as  a  boiler 
would.  If,  now/the  engine  starts  again,  steam  will  be  delivered 
to  the  accumulator  at  a  temperature  slightly  above  that  to  which 
the  water  has  fallen,  due  to  the  cooling  effect  of  the  evaporation 
for  supplying  the  turbine,  and  the  mass  of  water  will  again  absorb 


252 


THE  STEAM  TURBINE 


i-^CW 

]  C 
3  C 

"h                           jf 

]                          C 

3                         C 
3                         t 

3  "  'C 

J3                         411 

LOW-PRESSURE  STEAM  TURBINES  253 

heat  from  the  exhaust  steam.  Water  has  a  higher  specific  heat 
than  any  other  substance  except  hydrogen,  so  that  it  is  a  most 
suitable  and  convenient  substance  for  heat  accumulation. 

In  actual  practice  it  is  more  convenient  to  run  the  regenerator 
at  a  pound  or  two  pressure  above  the  atmosphere,  as  in  this  case 
the  piping  is  not  under  vacuum,  so  that  so  much  care  does  not 
have  to  be  exercised  to  avoid  air  leaks.  In  certain  cases,  how- 
ever, it  is  desirable  to  run  below  atmospheric  pressure.  In  this 
way  the  power  of  the  primary  engine  may  be  augmented  by 
letting  it  operate  at  a  partial  vacuum.  Plants  are  actually  run- 
ning with  a  delivery  pressure  to  the  turbines  as  low  as  six  pounds 
below  atmospheric  pressure. 

On  account  of  heat  radiation  from  the  accumulator,  water 
gradually  accumulates  in  it;  but  by  means  of  a  float  trap  shown 
at  the  right-hand  side  of  the  longitudinal  section  this  excess  of 
water  is  removed. 

If  for  any  reason  the  engine  shuts  down  for  a  considerable 
period,  the  supply  of  heat  stored  in  the  accumulator  will  become 
exhausted  and  the  pressure  will  fall  below  the  practical  limit  for 
operation  of  the  turbine.  To  provide  for  such  an  emergency  an 
automatic  reducing  valve  is  inserted  in  the  piping  to  deliver  live 
steam  to  the  accumulator.  There  is  also  a  relief  valve  on  the 
accumulator  through  which  excess  steam  will  pass  off  into  the 
air  when  the  pressure  becomes  3  or  4  pounds  above  atmospheric. 

The  pressure  in  the  accumulator  should  be  always  about  five 
to  ten  pounds  per  square  inch,  gage  pressure,  or  at  least  a  few 
pounds  above  the  atmospheric  to  avoid  the  possibility  of  air 
leaking  into  the  system. 

The  important  consideration  in  the  selection  and  designing  of 
an  accumulator  for  a  low-pressure  turbine  is  the  length  of  time 
the  regenerator  will  be  expected  to  carry  full  load  on  the  turbine 
without  receiving  any  low-pressure  steam  from  the  engine  or 
engines.  Obviously  the  longer  this  time  is,  the  greater  the 
capacity  required  of  the  accumulator.  Quite  generally  the  mis- 
take has  been  made,  according  to  Hodgkinson,*  of  supplying 

*  Tlie  Electric  Journal,  April,  1913,  page  335. 


254  THE   STEAM   TURBINE 

these  accumulators  in  much  too  large  sizes  for  the  require- 
ments. In  many  cases  the  time  interval  has  been  assumed  to 
be  six  to  seven  minutes,  during  which  the  accumulator  must 
supply  the  steam,  while  more  careful  study  shows  that  five  to 
six  seconds  would  have  been  a  much  better  estimate.  The  case 
of  a  steel  mill  is  cited.  If  the  exhaust  is  to  be  taken  from  a 
blooming  mill  the  time  element  should  bear  some  relation  to  the 
period  between  the  passes  of  an  ingot,  as  well  as  to  the  maximum 
tune  from  the  last  pass  of  one  ingot  to  the  first  pass  of  the  next 
ingot.  The  accumulator  should  not  be  designed,  therefore,  to 
cover  such  delays  as  would  arise  from  the  clogging  of  the  mills 
or  because  a  new  ingot  might  not  be  ready  to  be  bloomed.  For 
these  cases  of  unusual  delays  another  method  is  recommended. 
When  the  demand  for  steam  at  the  engines  is  interrupted  there 
will  be  a  sudden  rise  of  pressure  in  the  boilers  and  the  safety- 
valves  will  blow  off.  This  steam  should  be  piped  to  the  accu- 
mulator inlet,  instead  of  being  allowed  to  escape  to  the  atmos- 
phere. This  steam  from  the  safety  valves  will  assist  materially 
in  helping  the  low-pressure  turbine  in  carrying  its  load.  A  very 
good  arrangement  for  the  accomplishment  of  this  idea  is  to  place 
-a  "  cross-connection  "  of  piping  between  the  steam  main  sup- 
plying the  engine  and  the  engine  exhaust  line,  and  to  put  into 
this  line  a  globe  type  of  spring  loaded  valve  set  to  permit  steam 
to  pass  through  the  cross-connection  when  the  pressure  is  a  few 
pounds  lower  than  that  at  which  the  boiler  safety  valves  will 
blow. 

For  engine  power  plants,  where  the  supply  of  exhaust  steam 
is  often  stopped  for  long  periods,  the  accumulator  installation 
is  usually  dispensed  with,  and  the  low-pressure  turbines  are 
provided  with  piping  to  take  steam  directly  from  the  boilers, 
in  addition  to  the  exhaust  steam  piping.  (See  pages  258a  to 


An  exhaust  steam  turbine  has,  of  course,  relatively  few  rows  of 
blades  compared  with  ordinary  high-pressure  turbines. 

From  several  tests  made  with  5oo-kilowatt  exhaust  turbines 
in  England,  a  steam  consumption  of  34  pounds  per  kilowatt-hour 


LOW-PRESSURE   STEAM  TURBINES 


255 


was  obtained  with  15  pounds  per  square  inch  admission  pressure 
and  28  inches  vacuum. 

The  curve  in  Fig.  183  shows  the  steam  consumption  in  pounds 
per  horsepower-hour  at  the  switchboard  of  a  5oo-kilowatt 
exhaust  steam  turbine  of  the  Rateau  type. 


-f--+  -H--  Illllllllll  -- 

CURVE  OF  STEAM  CONSUMPTION 

H.P.-Hour  <it  Switchboard 

s  s  s 

MIN  |-[||fHJg 

RATEAU  STEAM  TURBINE 
-    INTERNATIONAL  HARVESTER  CO. 

.....,^._.._... 

Pounds  per 
8 

--•!;-  FJAHKai 

ttf  

300 


400 


600  600 

H.P.  at  Switchboard 


700 


FIG.  183.     Curve  of  Steam  Consumption  of  a  Rateau  Low-Pressure  Turbine. 

Some  tests  quoted  by  Francis  Hodgkinson  on  a  Westinghouse 
low-pressure  turbine  made  recently  gave  the  following  results : 


Steam  Pressure, 
Lbs.  per  Square 
Inch  Absolute, 
Dry  and  Satu- 
rated Steam. 

Vacuum  in  Exhaust  ,  ' 
Inches  Mercury  Re- 
ferred to  30  Inch 
Barometer. 

Load  in 
Brake  Horse- 
power. 

Total  Steam 
per  Hour. 

Steam  Con- 
sumption 
Brake  Horse- 
power Hour. 

17.4 
12.4 

25.98 
25.99 

920 
*    /  472 

25,670 
I7>487 

27.9 

37-i 

II.  8 

7-7 

5-2 

26.97 
27.03 
26.98 

592 
321 
IO2 

17,720 
11,980 

6>57° 

29.9 

37-3 
64-4 

ii.  6 

8.7 

27.8 
28.00 

586 

458 

16,400 
13,920 

28.0 
3°-4 

6.1 

4-5 

27.90 
27.99 

234 
114 

9.036 
6,248 

38.6 

54-8 

256 


THE   STEAM  TURBINE 


Fig.  184  is  a  copy  of  a  shop  drawing  of  a  looo-kilowatt  Westing- 
house  double-flow  low-pressure  turbine.  The  exhaust  steam  from 
the  engines  enters  through  the  annular  space  H  and  is  distributed 


FIG.     184.     looo-kilowatt-Westinghouse  Low-Pressure  Turbine. 

to  the  right  and  left  sections  of  Parsons  blading.  The  upper  half 
of  the  drawing  is  a  section  of  the  rotor  and  shows  the  method  of 
construction.  The  exhaust  is  discharged  through  the  base  as 
indicated  by  arrows.  The  openings  I,  I  are  provided  for  con- 
venient inspection  of  the  blading.  They  are  covered  with  suitable 
covers  in  which  automatic  relief  valves  are  fitted. 

Economy  curves  of  this  turbine  are  shown  in  Fig.  185.  The 
pressure  of  the  steam  delivered  to  the  turbine  was  approximately 
atmospheric.  The  vacuum,  as  shown  by  the  curves,  was  27^ 
inches  for  one  test  and  28  inches  for  the  other. 

Another  Westinghouse  turbine  built  to  operate  in  connection 
with  high-pressure  reciprocating  engines  gave  the  following 
results  in  a  shop  test: 

Initial  steam  pressure,  15  pounds  per  square  inch  absolute. 

Superheat,  40  degrees  F. 

Vacuum  referred  to  30  inch  barometer,  23  inches. 

Load,  1500  brake  horsepower. 

Steam  per  brake  horsepower  hour,  35.5  pounds. 


LOW-PRESSURE  STEAM  TURBINES 


257 


In  all  these  tests  the  exhaust  was  condensed  in  a  surface  con- 
denser, which  assures  accuracy  in  measuring  the  steam  con- 
sumption. 

A  Curtis  exhaust  steam  turbine  installed  in  Philadelphia 
receives  the  exhaust  steam  from  reciprocating  engines  at  a  pres- 
sure of  15  to  1 6  pounds  per  square  inch  absolute,  and  exhausts 
into  a  condenser  with  an  average  vacuum  of  28  inches.  The 
turbine  has  no  governor,  but  takes  all  the  steam  the  engines 
will  supply.  The  output  over  and  above  that  obtained  from 


•1-50 

-S43 


.5 


KO  K3  3CO  4< 


O  51  0  6(  0   *. 


""£ 


I?P1   1830 


IK o  ifl o  uwnwaxjfcoo  o  ia JOB x> K  10200 


FlG.  185.     Curves  of  Steam  Consumption  of  a  looo-kilowatt  Westinghouj 
Low-Pressure  Turbine. 


the  reciprocating  engines  is  increased  about  66  per  cent,  for  the 
same  steam  consumption.  With  dry,  saturated  steam  at  atmos- 
pheric pressure  delivered  to  the  turbine,  the  guaranteed  steam 
consumption  is  36  pounds  per  kilowatt  at  full  load,  and  40 
pounds  per  kilowatt  at  half  load.  It  is  stated  that  the  actual 
test  results  are  probably  at  least  10  per  cent,  better. 

Since  the  exhaust  from  reciprocating  engines  is  often  very  wet, 
it  is  good  practice  to  insert  a  steam  separator  in  the  steam  pipe 
leading  to  the  low-pressure  turbine. 

Most  applications  of  low-pressure  turbines  have  been  made  in 
collieries  and  steel  mills,  where  non-condensing  engines  are  the 
'rule.  Results  are,  however,  so  satisfactory  that  the  design  of 


258  THE  STEAM  TURBINE 

new  plants  having  a  compound  engine  with  a  smaller  low-pressure 
cylinder  than  is  usually  provided,  which  .is  to  discharge  its 
exhaust  into  a  steam  turbine  is  likely  to  become  common,  as 
giving  better  steam  economy  than  can  be  obtained  from  either 
reciprocating  engines  alone  or  turbines  alone. 

Low-Pressure  Steam  Turbines  Combined  with  Gas  Engines. 
There  is  also  another  field  open  to  the  low-pressure  steam  turbine. 
The  hot  cooling  water  from  the  jackets  of  large  gas  engines  could 
be  heated  by  the  exhaust  gases,  and  the  low-pressure  steam  thus 
formed  would  drive  a  steam  turbine. 


CHAPTER  X. 
MIXED-PRESSURE  TURBINES. 

IT  sometimes  happens  that  there  is  an  available  source  of  low- 
pressure  steam  which  it  is  desired  to  utilize  for  the  development 
of  power,  but  where  unfortunately  there  is  not  always  at  hand 
a  sufficiently  large  amount  of  this  low-pressure  steam  to  take 
care  of  the  power  requirement.  To  suit  this  condition  it  is  not 
infrequent  to  provide  in  a  steam  turbine  a  high-pressure  section 
to  take  steam  at  boiler  pressure  and  thus  help  out  the  low- 
pressure  section  when  its  normal  supply  of  steam  is  low.  In  this 
type  of  construction  when,  however,  the  supply  of  low-pressure 
steam  is  sufficient  for  the  power  requirements,  the  supply  of 
high-pressure  steam  is  cut  off  entirely  by  the  governor;  and  on 
the  other  hand  when  the  supply  of  low-pressure  steam  becomes 
again  sufficient  in  quantity,  no  more  high-pressure  steam  is 
used. 

The  generous  use  of  live  steam  in  low-pressure  steam  turbines 
is  not  by  any  means  as  poor  engineering  practice  as  at  first 
thought  it  appears.  The  obvious  reason  for  admitting  live  steam 
to  the  turbine  is  that  the  supply  of  low-pressure  steam  from 
the  engines  is  insufficient  for  the  turbine  requirements,  and  that 
consequently  some  of  the  engines  have  been  relieved  of  their 
load  more  or  less  suddenly.  The  boiler  plant  continues,  however, 
to  make  steam  at  the  former  rate,  and  the  safety  valves  will  soon 
blow  off  unless  the  excess  steam  can  be  used  in  the  power  plant. 
By  taking  this  excess  of  steam  to  the  turbine  to  help  in  carrying 
its  load,  which  we  shall  assume  has  not  been  reduced,  will  serve 
to  use  this  excess  of  steam  to  the  best  possible  advantage.  It 
is  not  an  unusual  practice  even  to  pipe  the  discharge  from  the 
safety  valves  on  the  boilers  into  the  receiver  in  the  low-pressure 
piping  supplying  the  turbine.  These  conditions  are  met  most 


258b  THE   STEAM  TURBINE 

frequently  in  the  suddenly  variable  loads  in  rolling  mills  and  in 
hoisting  operations  where  the  reciprocating  engine  drives  the 
rolls  or  hoists  and  the  low-pressure  turbine  supplies  a  more  or 
less  constant  electrical  load  for  both  lighting  and  comparatively 
light  power  requirements.  If  the  intervals  requiring  the  use 
of  high-pressure  steam  are  relatively  long,  as  for  example  five 
to  ten  hours  on  the  average,  then  a  so-called  "  mixed  turbine " 
type  should  be  used.  The  accumulator  method  is  also  adap- 
table for  the  longer  period  but  is  expensive  as  regards  first  cost. 
The  so-called  " mixed"  steam  turbine  has  been  a  development 
of  the  applications  of  steam  turbines  to  suit  two  important  con- 
ditions of  operation  which  are  as  follows:  (i)  the  case  where 
a  low-pressure  turbine  is  to  be  used  to  develop  an  amount  of 
power  for  which  there  is  not  constantly  available  a  sufficient 
amount  of  low-pressure  steam  to  carry  the  average  load;  and 
(2)  when  there  are  large  enough  quantities  of  low-pressure  steam 
at  certain  times  to  carry  the  load  but  at  more  or  less  long  inter- 
vals there  is  no  exhaust  steam  supplied  at  all.  Both  of  these 
cases  require  the  supplying  of  large  quantities  of  steam  from 
sources  independent  on  the  exhaust  lines  and  live  steam  direct 
from  the  boilers  is  invariably  the  substitute.  For  this  sort  of 
service  with  widely  varying  steam  pressures  the  mixed-pressure 
turbine  has  found  acceptable  application.  In  speaking  of  a 
mixed-pressure  turbine  in  this  chapter  we  shall  think  of  one  hav- 
ing separate  high-  and  low-pressure  portions  in  a  single  casing. 
A  good  example  is  shown  in  Fig.  1853,  where  the  high-pressure 
portion  is  provided  with  an  impulse  wheel  which  is  made  easily 
removable  so  that  when  for  long  periods  high-pressure  steam 
is*  not  needed  it  can  be  taken  off.  High-pressure  steam  enters 
at  the  steam  chest  opposite  the  nozzles  in  the  "impulse"  section. 
Low-pressure  steam  enters  only  the  reaction  blading  through  the 
vertical  pipe  coming  up  behind  the  " reaction"  section.  Such  a 
turbine  differs  essentially  from  the  ordinary  low-pressure  tur- 
bine which  is  provided  with  its  own  governor  only  in  having 
under  the  control  of  the  governor  a  special  set  of  valves  arranged 
to  supply  live  steam  to  nozzles  directing  steam  into  a  section  of 


MIXED-PRESSURE   TURBINES 


258C 


high-pressure  blades  before  discharging  through  the  low-pressure 
sections  along  with  a  supply  of  low-pressure  steam  with  which 
it  mixes.  A  very  common  type  of  mixed-pressure  turbine  con- 


FIG.  i85a.     Mixed-pressure  Turbine. 

sists  of  an  impulse  wheel  in  which  the  energy  drop  from  boiler 
to  about  atmospheric  pressure  is  absorbed  and  the  remainder 
of  the  energy  is  taken  out  by  expansion  in  low-pressure  reaction 
blading.  An  installation  of  this  kind  is  illustrated  by  Fig.  i8sb. 
The  valves  under  the  control  of  the  governor  are  adjusted  so 
that  no  high-pressure  steam  is  admitted  until  the  valves  on  the 
low-pressure  line  are  wide  open.  There  are  often  excellent  oppor- 
tunities for  the  installation  of  mixed-pressure  turbines  in  con- 
junction with  accumulators;  but  in  every  case  a  check  valve 
must  be  provided  between  the  accumulator  and  the  turbine  in 
the  low-pressure  line  to  prevent  live  steam  getting  back  into 
the  shell  of  the  accumulator,  which  will  probably  not  be  strong 
enough  to  withstand  the  excessive  stresses  that  might  be  pro- 
duced. 

In  many  mixed-pressure  turbines  the  high-pressure  section, 
if  of  a  simple  disk  construction,  is  frequently  made  removable 


258d 


THE   STEAM  TURBINE 


as  in  Fig.  i8sa,  so  that  the  " windage"  loss  due  to  its  revolution 
when  not  in  use  can  be  eliminated.  A  good  estimate  is  that 
about  2  per  cent,  of  the  power  of  the  turbine  is  lost  in  the  air 
resistance  of  such  a  high-pressure  section. 


FIG.  iSsb.     Low-pressure  Turbine  with  Live  Steam  Valve  Installed  at  Peace 

Dale,  R.  I. 

Mixed-pressure  turbines  are  not  often  used  in  sizes  larger 
than  about  2000  kilowatts.  To  meet  the  requirements  of  the 
larger  capacities  it  is  best  to  use  a  regular  or  "complete  expan- 
sion" turbine  in  combination  with  a  smaller  simple  low-pressure 
turbine. 


CHAPTER  XI. 
BLEEDER  OR  EXTRACTION  TURBINES. 

THE  name  " bleeder"  or  ''extraction"  turbine  is  given  to 
one  specially  designed  to  take  steam  at  boiler  pressure  and  to 
exhaust  part  of  this  steam  at  a  normally  low  vacuum  while 
another  part  is  "  extracted  "  or  taken  out  from  one  of  the 
stages  at  a  pressure  of  five  to  ten  pounds  per  square  inch  gage 
pressure;  that  is,  just  a  little  above  atmospheric.  In  many 
cases,  as  for  example  in  cotton,  woolen,  and  paper  mills,  this 
steam  is  " extracted"  for  manufacturing  purposes,  usually  heat- 
ing water  in  vats.  More  commonly,  however,  such  turbines 
find  their  application  for  supplying  the  low-pressure  steam  re- 
quired in  a  heating  system  for  houses,  factories,  office  buildings, 
etc.  Because  this  latter  supply  is  needed  only  a  part  of  the  year 
and  otherwise  is  variable  with  the  seasons,  there  will  be  times 
when  the  turbine  operates  by  complete  expansion  of  all  the  steam 
supplied  to  it  by  the  boilers.  Obviously  it  is  necessary  to  provide 
in  such  turbines  a  means  whereby  the  "  bleeder  "  steam  can  be 
taken  out  at  any  time,  and  with  sufficient  back  pressure  even  at 
light  loads  to  maintain  a  pressure  in  the  section  from  which  the 
steam  is  to  be  withdrawn  to  overcome  the  resistances  of  pipes 
and  valves,  so  that  steam  can  flow  freely  as  required.  It  is 
desirable  also  that  the  pressure  in  this  section  of  the  turbine 
should  be  fairly  constant.  To  accomplish  this  result  in  Wes- 
tinghouse-Parsons  turbines  a  partition  diaphragm  has  been  used 
to  separate  completely  the  high-pressure  from  the  low-pressure 
portion,  as  shown  in  Fig.  1850.  The  steam  from  the  high-pres- 
sure section  passes  normally  out  through  the  bleeder  passage 
and  into  the  mains  to  be  supplied  with  steam.  When,  how- 
ever, the  turbine  uses  more  steam  than  is  needed  for  the  service 
supplied  by  these  mains  then  the  pressure  in  these  passages 

2586 


THE   STEAM   TURBINE 


"  backs  up,"  and  when  above  that  pressure  for  which  the  auto- 
matic *  valve  shown  at  the  top  of  the  figure  has  been  set,  this 
steam  flows  out  into  the  low-pressure  sections  of  the  turbine, 
where  it  does  work,  and  thence  into  the  condenser.  It  is  the 
function  of  this  valve  to  maintain  sufficient  pressure  in  the 
passages  to  create  the  desired  flow  into  the  mains.  There  is 
therefore  some  throttling  action  in  this  valve  which  causes  a 
slight  loss  in  the  available  energy  of  the  steam  supplied  to  the 
low-pressure  sections. 


FIG.  1850.     Westinghouse-Parsons  "  Bleeder  "  Turbine. 

To  avoid  as  much  as  possible  the  throttling  action  referred 
to  above,  Curtis  turbines  are  designed  without  any  additional 
partitions  or  diaphragms.  The  method  adopted  is  to  place  a 
ring-shaped  valve  over  the  nozzles  leading  from  the  stage  from 
which  the  steam  is  to  be  "  extracted.1'  This  valve  is  operated 
automatically  by  a  mechanism  responsive  to  the  pressure  in  the 
stage  so  that  the  effective  area  of  the  nozzles  is  changed  as 
required  to  maintain  a  constant  pressure.  By  this  method  the 
closing  of  the  nozzles  occurs  only  in  groups,  so  that  any  slight 
throttling  action  that  might  occur  due  to  partial  opening  would 
create  its  loss  only  in  a  small  group  rather  than  in  all  the  nozzles 

*  Similar  to  a  relief  or  safety  valve  in  its  action. 


BLEEDER  OR  EXTRACTION  TURBINES 


in  the  stage;  those  not  throttled  would  be  either  closed  off 
entirely  or  else  wide  open. 


FIG.  i85d.     Ring  Valve  of  Curtis  Bleeder  Turbine. 


FIG.  1856.     Side  of  Diaphragm  of  Bleeder  Turbine. 

Parts  of  this  device  are  illustrated  by  the  following  figures: 
Fig.  i8sd  shows  the  ring  valve  used  for  covering  the  nozzles 


THE   STEAM  TURBINE 


which  are  bolted  in  the  usual  construction  to  the  side  of  the 
diaphragm  (Fig.  1856).  This  valve  is  operated  by  the  piston  in 
an  oil  (or  steam)  cylinder  which  is  in  turn  moved  by  being 
subjected  to  oil  or  (steam)  under  pressure  admitted  from  a  high- 
pressure  supply  by  a  small  pilot  valve  actuated  by  leverage  con- 
nections to  the  diaphragm  (30)  in  communication  by  small  piping 
with  the  stage  in  which  the  constant  pressure  is  to  be  main- 
tained. Fig.  i8sf  shows  a  cross-sectional  view  of  the  mechan- 


/  I 

FIG.  185!.    Valve  Gear  of  Curtis  Bleeder  Turbine. 

ism  which  actuates  the  valve.  By  means  of  a  flexible  joint  at 
(i)  the  piston  rod  (2)  moves  the  valve  plate  back  and  forth  over 
the  face  of  the  nozzles.  Pressure  on  the  piston  (14)  in  the  oil 
(or  steam)  cylinder  (13)  gives  the  movement  to  the  piston  rod. 
Movements  of  the  piston  are  effected  by  means  of  the  pilot  valve 
(22)  which  is  in  turn  actuated  by  the  diaphragm  (30)  by  means 
of  the  rods  (34)  and  (39) .  This  diaphragm  with  its  "  corrugated  " 
or  "  accordion  "  sides  forms  a  cylindrical  chamber  which  is  in 
communication  by  means  of  small  piping  with  the  stage  to  be  con- 
trolled, and  from  which  the  "  bleeder  "  steam  is  to  be  taken. 


BLEEDER   OR   EXTRACTION  TURBINES  2581 

Movements  of  the  diaphragm  are  opposed  by  the  spiral  spring 
(36)  which  can  be  set  to  maintain  any  desired  steam  pressure  in 
the  stage. 

Careful  inspection  of  Fig.  i8sd  shows  that  the  ports  in  the 
ring  valve  are  not  all  of  the  same  size  but  are  of  progressively 
increasing  width  around  the  circumference  from  the  narrowest 
to  the  largest.  The  narrow  parts  begin  closing  up  on  the  first 
movement  of  the  valve.  There  are  four  groups.  The  second 
group  begins  closing  only  after  the  first  or  narrowest  set  is 
fully  closed. 

A  balance-plate  (Fig.  i8sg)  is  put  on  top  of  the  ring  valve 
(Fig.  i8sd)  for  the  purpose  of  assisting  in  equalizing  the  pres- 
sure on  the  two  sides  of  the  valve,  and  thus  reducing  the  force 
required  to  move  it. 


FIG.  i8sg.      Nozzle -plate  for  Curtis  Ring  Valve. 

It  is  comparatively  a  very  easy  matter  to  remove  some  of  the 
steam  which  has  been  partly  expanded  in  the  turbine  by  the 
use  of  suitable  automatic  or  hand-controlled  valves  even  when 
the  quantity  of  steam  required  at  a  constant  pressure  in  such  a 
bleeder  line  is  quite  variable.  By  this  method  it  is  possible  to 


258J 


THE   STEAM   TURBINE 


extract  the  greatest  amount  normally  possible  as  required  for 
generating  power  and  at  the  same  time  supplying  at  a  reason- 
ably constant  pressure  "usually  about  atmospheric,"  or  about 
5  pounds  above,  the  requirements  for  heating  or  industrial 
purposes. 

Fig.  i8sh  shows  the  satisfactory  filling  of  the  blades  of  an 
impulse  turbine  of  the  "  bleeder  "  type. 


FIG.  i8sh.     Flow  Lines  in  a  "  Bleeder  "  Impulse  Turbine. 


CHAPTER  XII. 
MARINE  TURBINES, 

ONE  of  the  most  important  fields  for  the  steam  turbine  is  the 
propulsion  of  ships.  In  the  mercantile  marine  the  progress  of 
the  turbine  had  been  extremely  rapid,  the  first  mercantile  vessel 
propelled  by  turbines  having  been  built  only  a  very  few  years 
ago.  That  vessel  had  about  700  tons  displacement,  and  developed 
3500  indicated  horsepower,  comparing  with  a  tonnage  of  45,000 
and  70,000  horsepower  in  the  Lusitania  and  Mauretania.  Care- 
ful trials  had  shown  that  at  all  speeds  above  14  knots  the  turbine 
was  more  economical  than  the  reciprocating  engine,  being  15  per 
cent,  better  at  18  knots,  31  per  cent,  better  at  20  J  knots,  and  36 
per  cent,  better  at  20.  i  knots.  In  the  Dover-Calais  service  it  had 
been  found  that  the  turbine  boats  carried  passengers  at  two  knots 
greater  speed  with  25  per  cento  less  coal  per  passenger  than  boats 
propelled  with  reciprocating  engines.  A  saving  in  coaj  of  about 
9  per  cent,  was  computed  for  the  turbine  steamers  belonging  to  the 
Midland  Railway  (England),  as  compared  with  similar  steamers 
of  the  same  company  equipped  with  reciprocating  engines. 
The  difference  in  initial  cost  and  in  weight  of  machinery  was 
found  to  favor  the  turbine  driven  ships  by  ij  and  6  per  cent, 
respectively.  When  used  for  marine  service,  doubtless  the 
greatest  defect  of  practical  steam  turbines  is  that  they  cannot  be 
reversed.  Many  attempts  have  been  made  to  devise  a  turbine  to 
reverse  in  a  simple  way  comparable  with  a  reversing  reciprocat- 
ing engine.  It  is  the  present  practice  to  provide  turbine  driven 
ships  with  two  turbines  used  only  for  reversing,  and  as  they  are  not 
intended  for  high  speed,  they  may  be  of  small  power  compared 
with  the  main  turbines.  These  two  reversing  turbines  are  usually 
fitted  to  the  same  shafts  as  the  low-pressure  turbines,  and  when 
the  ship  is  running  ahead  their  rotors  revolve  idly  in  a  vacuum. 

259 


260  THE   STEAM  TURBINE 

When  the  ship  is  to  be  run  backward  the  steam  is  shut  off  from 
the  "ahead"  turbines  and  is  admitted  to  the  auxiliary  reversing 
turbines.  There  is,  of  course,  a  disadvantage  from  not  having 
at  times  the  full  normal  motive  power  of  the  ship  available 
for  backing.  Besides,  conditions  are  not  ideal  when  a  large 
portion  of  the  plant  is  idle  for  a  greater  part  of  the  time.  These 
reversing  turbines  will  occupy  a  great  deal  of  longitudinal  space, 
so  that  the  floor  space  required  for  an  installation  of  marine 
steam  turbines  is  larger  than  that  required  for  reciprocating 
engines  for  the  same  conditions  of  service. 

The  White  Star  Company  (International  Mercantile  Marine 
Company)  has  decided  to  operate  ocean  steamers  with  a  com- 
bined reciprocating  and  turbine  engine  plant.  The  two  outer 
shafts  will  be  driven  by  quadruple  expansion  reciprocating  en- 
gines and  the  central  shaft  by  a  low-pressure  turbine  operated 
by  the  exhaust  steam  from  the  low-pressure  cylinder  of  the 
reciprocating  engines.  For  going  backward,  the  reciprocating 
engines  will  be  used,  as  they  are  readily  reversed,  and  in  the 
ordinary  service  the  turbine  and  reciprocating  engines  will  be 
operated  together.  By  this  combination  the  advantages  of 
reciprocating  engines  for  reversing  are  secured,  together  with 
the  great  range  of  expansion  which  is  possible  with  the  steam 
turbine. 

It  is  difficult  to  say  what  developments  the  future  will  bring 
in  the  applications  of  steam  and  gas  turbines  to  the  marine 
service.  Practically  all  the  new  battleships  and  cruisers  for  the 
British  navy  are  now  turbine  driven.  If  we  consider  that  the 
steam  turbine  in  its  practical  form  commenced  its  real  develop- 
ment only  in  1885,  the  future  certainly  may  have  rich  possibilities. 

Fig.  185!  represents  the  results  of  tests  made  at  variable 
speeds  and  powers  on  a  standard  combined  impulse  and  reaction 
type  of  turbine.  In  explaining  the  results  of  these  tests  Mr. 
H.  T.  Herr  *  states  that  investigations  now  under  way  by  the 
"  Westinghouse  interests  "  will  insure  the  elimination  in  the  near 
future  of  the  reciprocating  engine  in  the  field  of  marine  propul- 
*  Journal  of  the  Franklin  Institute,  March,  1913. 


MARINE   TURBINES 


26oa 


sion  as  the  turbine  generator  has  practically  eliminated  it  for 
electric  power  plant  service. 

On  account  of  the  difficulty  of  adjusting  the  inherent  re- 
quirements of  the  steam  turbine  for  operation  at  relatively  high 
rotative  speed  and  the  corresponding  difficulty,  opposite  how- 
ever, in  effect,  of  the  efficient  operation  of  the  propellers  of 
steamships  at  anything  but  relatively  low  speed,  the  applica- 
tions of  steam  turbines  to  the  propulsion  of  ships  has  been  very 
much  limited.  If  it  were  not  for  these  difficulties  there  is  no 
reason  why  the  steam  turbine  should  not  displace  the  recipro- 


-s 

^   § 

a  rf» 


45 

40 

30    - 

o  S  g  8  £  §  §  2 

Efficiency  in  per  cent. 

O  i  i  ~-  ~.  s.  ~  t±  *-  -.  j 

Pounds  of  steam  per  brake  horsepower  hour. 

\. 

1 

\ 

i 

\ 

/ 

s4 

s 

^ 

7 

S 

^ 

r—  l 

—  -- 

s 

^ 

^^< 

•** 

/ 

?„•>     . 

><j 

X 

^^ 

L. 

j 

20    - 

/ 

, 

^ 

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j 

^ 

V 

$ 

j 

x' 

*/> 

/ 

10 

< 

/ 

^ 

*^ 

/ 

ev 

/ 

^f 

V 

2 

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5    - 

/ 

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x 

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^^ 

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2.5 

/ 

^-- 

,  —  • 

^ 

^ 

2 

f  — 

—  ' 

800 


1600 


2400          3200         4000 
Revolutions  per  minute. 


FIG.  1851. 


Curves  Showing  Variation  of  Steam  Consumption,  Horsepower 
Efficiency  of  Latest  Designs  of  Steam  Turbines  with  Speed. 


,  and 


eating  steam  engine  almost  entirely  for  this  service.  On  this 
account,  however,  the  application  has  been  confined  almost 
entirely  to  merchant  and  naval  vessels  designed  for  high-speed 
service.  Experience  has  shown  that  in  the  applications  of  steam 
turbines  to  slow-speed  ships  there  has  been  no  appreciable 
saving  in  weight,  in  space,  or  in  the  cost  of  operation,  over  what 
it  would  have  been  with  reciprocating  engines.  The  nearest 
approach  to  the  solution  of  this  problem  is  to  be  secured  prob- 
ably by  the  application  of  gearing  essentially  similar  to  that 
designed  by  De  Laval.  This  sort  of  gearing  cannot,  however, 


26ob  THE    STEAM   TURBINE 

be  applied  without  modification  to  turbines  developing  more 
than  possibly  1000  horsepower.  A  design  much  better  suited  to 
high-speed  conditions  and  also  adaptable  for  large  power  has 
been  developed  by  the  Westinghouse  Machine  Company  with 
the  cooperation  of  Mr.  George  Westinghouse,  Admiral  George 
Mellville,  and  Mr.  John  H.  Macalpine.  The  essential  principle 
embodied  in  their  improvements  consists  in  the  application  of  a 
so-called  "  floating  frame  "  designed  to  carry  the  pinion  on  the 
main  turbine  shaft.  The  experimental  gear  developed  in  the 
early  stages  had  a  floating  frame  supported  on  pivots,  permitting 
flexibility  as  regards  horizontal  movement  of  the  pinion,  but 
was  rigid  as  regards  vertical  movements.  Very  recently  Mr. 
Westinghouse  developed  a  very  important  improvement  con- 
sisting in  the  substitution  of  a  flexible  support  by  means  of 
hydraulic  pistons  taking  the  place  of  the  rigid  vertical  supports, 
and  in  this  way  improving  very  much  the  efficiency  and  the 
wearing  properties  of  the  gear.  This  improvement  in  wearing 
properties  had  also  the  effect  of  reducing  to  a  minimum  all 
noise  and  vibration  which  in  the  original  design  were  consid- 
erable. In  this  later  design  (Figs.  1853*  and  k),  the  main  frame 
supporting  the  pinion  is  held  up  by  the  pistons  in  the  hydraulic 
cylinders  filled  with  oil  under  pressure.  This  construction 
permits  vertical  movements  of  the  pinion  along  with  its  flexi- 
bility in  its  floating  frame  and  is  therefore  a  great  improve- 
ment in  that  the  earlier  design  permitted  only  lateral  move- 
ment. The  vertical  movement  is  permitted  by  the  supporting 
of  the  floating  frame  on  the  piston  connected  to  the  support- 
ing rods.  Similarly  the  lateral  movement  is  permitted  by  the 
flexibility  of  the  horizontal  pistons  on  the  two  sides  of  the 
pinion. 

For  a  more  complete  description  and  discussion  of  the  "  float- 
ing frame  "  type  of  reduction  gear  see  Engineering,  vol.  95  (1913)  > 
pages  169  and  609,  and  The  Electric  Journal,  January,  1912. 

Water  rate  curves  drawn  from  the  data  of  acceptance  tests 
of  the  battleships  North  Dakota  and  Delaware  are  shown  in 
Fig.  185!.  Curves  A  and  BI  show  the  steam  consumption  per 


MARINE  TURBINES 


2600 


FIG.  185].     "Floating  Frame"  Reduction  Gear,  Showing  Gears  when  Side  of 

Casing  is  Removed. 


26od 


THE   STEAM  TURBINE 


FIG.  iSsk.    Reduction  Gear,  Showing  Flexible  Support  of  Pinion. 


I18 

16 


N  12  04 


2000   6000   10000   14000   18000   22000   26000   30000 
Horsepower. 

FIG.  185!.     Water  Rate  Curves  of  U.  S.  Battleships  and  Computed  Curves  if 
Geared  Steam  Turbines  were  used. 


MARINE  TURBINES  26oe 

shaft  horsepower  per  hour  of  the  engines  in  the  two  battleships, 
while  C  and  Ci  show  the  corresponding  results  if  geared  turbines 
had  been  used  instead  of  reciprocating  engines.  It  is  estimated 
that  with  a  geared  turbine  combination  of  the  Westinghouse 
"  floating  frame  "  type  the  economy  of  the  prime  movers  in 
vessels  of  the  Delaware  class  could  be  improved  30  per  cent, 
at  full  speed  and  25  per  cent,  at  cruising  speed.  Fig.  i8sm 


FIG.  185111.    Low-pressure  Turbine  Casing  for  German  Steamship. 

shows  the  enormous  size  of  the  casings  for  the  low-pressure 
sections  of  the  steam  turbines  installed  in  modern  battle- 
ships. 

Electrical  Transmission  for  Ships.  Another  method  differ- 
ent from  the  use  of  reduction  gears  has  been  frequently  sug- 
gested for  making  the  steam  turbine  more  adaptable  for  marine 
service.  This  method  consists  in  using  on  the  vessel  steam 
turbines  direct  connected  to  high-speed  electric  generators, 
which  can  operate  then  under  practically  identical  conditions 
as  in  "  land  "  service.  These  turbines,  obviously,  can  then  be 
designed  to  operate  at  a  speed  best  suited  to  obtain  high  effi- 
ciency. The  electric  current  from  the  generators  is  used  to 
drive  slow-speed  electric  motors  on  the  shafts  of  the  propellers, 
the  speed  here  being  that  giving  best  efficiency  for  the  pro- 
pellers. 

This  method  offers  great  flexibility  in  the  handling  of  a 
vessel,  as  the  motors  can  be  very  quickly  reversed,  and  changes 
of  speed  are  readily  obtainable. 


26of  THE   STEAM   TURBINE 

Although  this  method  for  marine  propulsion  has  been  ad- 
vocated by  engineers  for  many  years  it  has  not  as  yet  re- 
ceived very  favorable  acceptance;  but,  doubtless,  it  is  con- 
stantly receiving  more  favorable  attention  from  well-known 
designers. 


CHAPTER   XIII. 
TESTS  OF  STEAM  TURBINES. 

Testing  Steam  Turbines.*  In  every  power  plant  the  means 
should  always  be  available  for  making  tests  of  the  steam  equip- 
ment to  determine  the  steam  consumption.  Usually  tests  are 
made  to  determine  how  nearly  the  performance  of  a  turbine 
approaches  the  conditions  for  which  it  was  designed.  The 
results  obtained  from  tests  of  a  turbine  are  to  show  usually  the 
steam  consumption  required  to  develop  a  unit  of  power  in  a  unit 
of  time,  as,  for  example,  a  horsepower-hour  or  a  kilowatt-hour. 

In  such  tests  a  number  of  observations  must  be  made  regarding 
the  condition  of  the  steam  in  its  passage  through  the  turbine  and 
of  the  performance  of  the  turbine  as  a  machine.  To  get  a  good 
idea  of  what  these  observations  mean,  it  may  be  profitable  to 
follow  the  steam  as  it  passes  through  the  turbine.  The  steam 
comes  from  the  boilers  through  the  main  steam  pipe  and  the 
valves  of  the  turbine  to  the  nozzles  or  stationary  blades  as  the 
case  may  be.  It  then  passes  through  the  blades  and  finally 
escapes  through  the  exhaust  pipe  to  the  condenser.  It  is  pref- 
erable to  have  a  surface  condenser  for  tests  so  that  the  exhaust 
steam  can  be  weighed.  The  weighing  is  done  usually  in  large 
tanks  mounted  on  platform  scales. 

Methods  for  Testing.  The  important  observations  to  be  made 
in  steam  turbine  tests  are: 

1.  Pressure  of  the  steam  supplied  to  the  turbine. 

2.  Speed  of  rotation  of  the  turbine  shaft,  usually  taken  in 
revolutions  per  minute. 

3.  Measurement  of  power  with  a  Prony  or  a  water  brake,  if 
the  power  at  the  turbine  shaft  is  desired ;  or  with  electrical  instru- 
ments (ammeters,  voltmeters,  and  wattmeters),  if  the  power  is 
measured  by  the  output  of  an  electric  generator. 

*  For  complete  and  detailed  information  regarding  the  testing  of  steam  tur- 
bines and  other  prime  movers,  as  well  as  the  revised  codes  of  testing  adopted  by 
the  American  Society  of  Mechanical  Engineers, .  see  Power  Plant  Testing,  pages 
294-363,  by  the  author  (McGraw-Hill  Book  Co.,  N.Y.). 

261 


262  THE  STEAM  TURBINE 

4.  Weight,   or  measurement   by  volume,   of  the   condensed 
steam  discharged  from  the  condenser.     Unless  a  surface  con- 
denser is  used  it  is  very  difficult  to  obtain  the  amount  of  steam 
used   by  the  turbine.     All  leakages  from   pipes,   pumps,   and 
valves,  which  is  part  of  the  steam  which  has  gone  through  the 
turbine,  must  be  added  to  the  weight  of  the  condensed  steam. 
The  accuracy  of  a  test  often  depends  a  great  deal  on  how 
accurately  leaks  have  been  provided  against,  or  measured  when 
they  occur. 

5.  Temperature  of  the  steam  as  it  enters  the  turbine.     If  the 
temperature  is  higher  than  that  due  to  the  pressure  of  the  saturated 
steam  given  in  steam  tables,  the  steam  is  superheated;  if,  how- 
ever, the  temperature  is  not  higher  the  steam  may  be  wet,  and  a 
calorimeter  must  be  attached  as  near  the  turbine  steam  chest  as 
possible.* 

All  gauges,  electrical  instruments,  and  thermometers  should  be 
carefully  calibrated  before  and  after  each  test  so  that  observations 
can  be  corrected  for  any  errors.  The  zero  readings  of  Prony 
and  water  brakes  for  measuring  power  should  be  carefully 
observed  and  corrected  to  eliminate  the  friction  of  the  apparatus 
with  no  load.  Unless  all  these  precautions  are  taken  the  dif- 
ficulties in  getting  reliable  tests  of  turbines  are  greatly  increased. 
In  all  cases  tests  should  be  continued  for  several  hours  with 
absolutely  constant  conditions  if  the  tests  are  to  be  of  value. 

The  most  valuable  test  of  a  steam  turbine  is  made  when  varying 
only  the  load;  that  is,  with  pressures,  superheat,  and  speed  con- 
stant. When  the  steam  consumption  is  then  plotted  against 
fractions  of  full  load,  a  water-rate  curve  is  obtained.  For  such 
a  curve  a  series  of  tests  are  needed,  each  for  some  fraction  of  full 
load ;  and  in  each  separate  test  the  power  as  well  as  all  the  other 
conditions  must  be  held  constant. 

*  The  most  satisfactory  tests  of  turbines  are  made  with  steam  slightly  super- 
heated rather  than  wet.  When  steam  is  very  wet  (more  than  about  4  per  cent, 
moisture  for  ordinary  pressures)  the  determination  of  the  quality  is  difficult.  There 
is  also  a  danger  that  steam  showing  only  a  few  degrees  of  superheat  by  the  reading 
of  the  thermometer  is  actually  wet.  The  high  temperature  is  due  in  such  cases  to 
heating  from  eddies  around  the  thermometer  case  or  in  steam  pockets  near  it. 


TESTS  OF   STEAM  TURBINES  263 

Another  important  test  of  the  performance  of  steam  turbines  is 
made  by  varying  both  the  speed  and  the  power  and  keeping  the 
other  conditions  constant.  The  observations  of  speed  and  power 
from  such  a  test  give  a  power  parabola  as  illustrated  in  Fig.  80. 
This  curve  shows  at  what  speed  the  turbine  gives  the  greatest 
output. 

For  complete  tests  of  a  steam  turbine  the  steam  consumption 
should  be  determined  at  full  load  (i)  with  varying  initial  steam 
pressure;  (2)  with  varying  vacuum;  and  (3)  with  varying  superheat. 

A  complete  set  of  tests  as  outlined  will  give  sufficient  data  to 
determine  all  the  corrections  usually  required. 

Commercial  Testing.  The  methods  used  by  the  New  York 
Edison  Company  in  commercial  tests  of  steam  turbine-generator 
units  may  well  be  explained  briefly. 

During  a  test  the  load  on  the  turbine  unit  is  maintained  as  con- 
stant as  possible  by  "remote  control"  of  the  turbine  governor  by 
the  switchboard  operator.  The  maximum  variation  in  load  is 
to  be  held  within  4  per  cent,  above  and  below  the  mean.  For 
some  time  previous  to  the  test  the  turbine  is  run  a  little  below  the 
load  required  for  the  test,  but  at  least  ten  minutes  before  the 
starting  signal  is  given  the  test  load  must  be  on  the  machine. 

Three-phase  electrical  load  is  measured  by  the  two-wattmeter 
method,*  using  Weston  indicating  wattmeters  of  the  standard 
laboratory  type.  These  instruments  are  calibrated  by  a  well- 
known  testing  laboratory  immediately  before  and  after  the  test. 
Power  factor  is  maintained  substantially  at  unity  and  all  electrical 
readings  are  taken  at  one-minute  intervals. 

When  the  turbine  is  provided  with  a  surface  condenser,  the 
steam  consumption,  or  water  rate,  is  determined  by  weighing  in 
a  large  tank  supported  on  platform  scales  the  condensed  steam 
delivered  from  the  condenser  hot  well.  Above  the  weighing 
tank  a  reservoir  is  provided  which  is  large  enough  to  hold  the 
condensation  accumulating  between  the  weighings  which  are 
made  at  intervals  of  five  minutes.  By  using  a  loop  connection 

*  Cf.  Kent's  Mechanical  Engineer's  Pocket-Book,  7th  ed.,  page  1069,  8th  ed., 
page  1396,  or  Foster's  Electrical  Engineer's  Pocket-Book. 


264  THE   STEAM  TURBINE 

for  the  gland  water  supply  (of  Westinghouse  turbines)  or  the 
water  from  the  step  bearing  (of  Curtis  turbines  using  water  for 
this  bearing)  the  necessity  for  correcting  the  weighings  for  these 
amounts  is  avoided. 

Because  the  circulating  water  at  the  stations  of  this  company 
is  usually  quite  salt,  any  condenser  leakage  is  detected  by  testing 
the  condensed  steam  by  the  silver-nitrate  method  with  a  suitable 
color  indicator.  This  color  method  is  said  to  be  a  decided 
advantage  over  the  usual  method  of  weighing  the  leakage  accu- 
mulating during  a  definite  period  when  the  condenser  is  idle  and 
is  tested  for  only  one  particular  vacuum.  By  taking  samples 
of  circulating  water  and  condensed  steam  at  the  same  time, 
it  is  possible  to  detect  any  change  in  the  rate  of  condenser 
leakage. 

The  water  level  in  the  hot  well  is  maintained  at  practically  a 
constant  point  by  means  of  a  float  valve  in  the  well  automatically 
controlling  the  speed  and,  therefore,  the  amount  of  the  delivery 
of  the  hot-well  pump.  This  device  avoids  the  necessity  for  the 
difficult  correction  to  be  made  in  a  test  when  the  levels  in  the  hot 
well  are  not  the  same  at  the  beginning  and  end  of  a  test.  Tem- 
peratures and  pressures  of  the  admission  steam  are  determined 
by  mercury  thermometers  and  pressure  gauges  located  near  the 
main  throttle  valve  of  the  turbine;  the  amount  of  superheat  is 
determined  by  subtracting  from  the  actual  steam  temperature 
after  making  thermometer  corrections  the  temperature  of 
saturated  steam  corresponding  to  the  pressure  at  the  point  where 
the  temperature  is  measured.  All  gauges  and  thermometers  are 
calibrated  before  and  after  the  test. 

Vacuum  is  measured  directly  at  the  turbine  exhaust  by  means 
of  a  mercury  column  with  a  barometer  alongside  for  reducing  the 
vacuum  to  standard  barometer  conditions  (30  inches).  By  this 
latter  arrangement  the  necessity  for  temperature  corrections 
which  are  necessary  when  the  two  mercury  columns  are  not  at 
the  same  place  is  avoided. 

Fig.  1 88  shows  a  55oo-kilowatt  Westinghouse-Parsons  turbine 
set  up  for  testing  in  the  shops  before  shipment  to  the  customer. 


TESTS  OF  STEAM  TURBINES 


265 


266 


THE   STEAM  TURBINE 


The  power  is  measured  by  means  of  a  large  water  brake  shown  in 
the  figure  at  the  left  of  the  turbine. 

Reports  of  Tests.  The  tables  given  below  have  been  prepared 
to  show  the  steam  consumption,  together  with  the  most  impor- 
tant other  data,  of  what  are  believed  to  be  reliable  tests  of 
standard  makes  of  steam  turbines.  The  vacuum  given  in  the 
tables  is  the  equivalent  referred  to  30  inches  barometer. 

Curtis  Turbines.  The  following  results  were  obtained  in 
1905  by  Messrs.  Sargent  and  Lundy  with  a  2ooo-kilowatt  Curtis 
turbine-generator. 


Kilowatts. 

Steam  Pressure 
(Gauge)  . 

Superheat, 
Deg.  F. 

Vacuum,  Inches. 

Pounds  per 
Kilowatt-hour. 

555 
1067 
2024 

155-5 
170.2 
166.3 

204 

120 
207 

28-5 
28.4 

28-5 

18.09 
16.31 
15.02 

Also  the  following  results  are  reported  in  1907  with  a  9000- 
kilowatt  turbine-generator  in  Chicago : 


Kilowatts. 

Steam  Pressure 
(Gauge). 

Superheat, 
Deg.  F. 

Vacuum, 
Inches. 

Pounds  per 
Kilowatt-hour. 

5.374 

182 

133 

29-43 

I3-I5 

8,070 

179 

116 

29-35 

13.00 

10,186 

176 

147 

29.47 

12.90 

13,900 

198 

140 

29.31 

13.60 

Parsons    Turbines.      A    i5oo-kilowatt    Parsons    turbine   was 
tested  at  Sheffield,  England,  with  the  following  results: 


Kilowatts. 

Steam  Pressure 
(Gauge). 

Superheat, 
Deg.  F. 

Vacuum, 
Inches. 

Pounds  per 
Kilowatt-hour. 

53° 
1071 

1585 

145.0 
131.0 
128.5 

no 
124 
125 

28.9 
28.3 

27-5 

21.58 
18.24 
17.60 

TESTS  OF  STEAM  TURBINES 


267 


The  results  of  two  tests  of  a  3oo-kilowatt  Parsons  turbine 
installed  at  the  Hulton  colliery  are  also  given  to  show  the  change 
of  economy  from  running  condensing  26.58  inches  vacuum  and 
non-condensing. 


Kilowatts. 

Steam  Pressure 
(Gauge). 

Superheat, 
Deg.  F. 

Vacuum, 
Inches. 

Pounds  per 
Kilowatt-hour. 

303 
297 

158.0 
i6i  .0 

0 

o 

26.6 

O. 

23-l5 
34-20 

These  last  tests  show  well  the  increased  steam  consumption 
(about  50  per  cent.)  when  running  non-condensing. 

Westinghouse-Parsons  Turbines.  The  table  below  gives  the 
results  of  tests  in  1904  by  F.  P.  Sheldon  &  Co.,  Providence, 
R.I.,  of  a  4oo-kilowatt  Westinghouse-Parsons  turbine  with  about 
100  degrees  F.  superheat. 


Brake 

Steam  Pressure 

Superheat, 

Vacuum, 

Pounds  per 

Horsepower. 

(Gauge). 

Deg.  F. 

Inches. 

B.H.P.  Hour.* 

279.4 

I53-I 

92-5 

28.0 

14-34 

410.7 
657-3 

153-2 
152.7 

102.9 
100.3 

28.0 
28.0 

.13-45 
12.48 

967.5 

149.6 

IOO.  2 

27.6 

12.79 

1207.5 

152.0 

99-9 

27-3 

13-55 

*  Observe  the  steam  consumption  is  in  pounds  per  brake  horsepower  hour,  instead  of  pounds  per 
kilowatt-hour  as  for  some  of  the  other  results  given  here. 

The  curves  given  in  Fig.  189  were  plotted  to  show  graphi- 
cally the  steam  consumption  of  300,  500,  and  1000  kilowatt 
Westinghouse-Parsons  turbines  with  varying  loads.*  Data  of 
the  tests  from  which  these  curves  were  drawn,  as  well  as  of  a 
test  of  a  3ooo-kilowatt  turbine  are  given  in  the  following  tables. 
These  tests  were  reported  by  J.  R.  Bibbins  in  1906  and  1907. 

*  The  numbers  marked  on  the  curves  to  indicate  the  vacuum  represent  the 
actual  readings  taken  in  the  test  and  are  not  referred  to  a  standard  (30  inches) 
barometer. 


268 


THE   STEAM  TURBINE 

m 


300-KILOWATT  TURBINE  (3600  R.P.M.). 


1    o 


u 


£ 

II 


U    «» 

I  £ 

.  <u 

£  8^ 

'o 

t/3 
I 


Brake 

Steam  Pressure 

Superheat, 

Vacuum, 

Pounds  per 

Horsepower. 

(Gauge). 

Deg.  F. 

Inches. 

B.H.P.  Hour. 

-232.7 
460.6 
688.5      ; 

I45-I 
144.6 
140.3 

4.1 
4.8 
7.0 

28.0 
28.0 

27.2 

15-99 
13-99 
15-73 

TESTS  OF  STEAM  TURBINES  269 

500-KILOWATT  TURBINE  (CONDENSING  AND  NON-CONDENSING)  (3600  R.P.M.). 


Brake 

Steam  Pressure 

Superheat, 

Vacuum, 

Pounds  per 

Horsepower. 

(Gauge). 

Deg.  F. 

Inches. 

B.H.P.  Hour. 

383.5 

152.6 

.2 

28.2 

I4-I5 

755-6 

149.2 

1.2 

27.8 

13.28 

1121.9 

148.8 

5-i 

26.5 

14.32 

385-6 

148.2 

2.7 

0.8 

24.94 

766.8 

147-3 

2.6 

0.8 

22.  10 

1144.4 

126.  i 

ii.  4 

0.8 

24.36 

These  last  tests  show  well  the  increased  steam  consumption 
(about  75  per  cent.)  when  running  non-condensing. 

1000-KILOWATT  TURBINE  (1800  R.P.M.). 


Brake 

Steam  Pressure 

Superheat, 

Vacuum, 

Pounds  per 

Horsepower. 

(Gauge). 

Deg.  F. 

Inches. 

B.H.P.  Hour. 

752.4 

JSo-S 

0.2 

27-5 

14.77 

I5°3  -5 

146.7 

0.0 

27.0 

13.61 

2252.7 

T45-3 

0.0 

25.2 

15.29 

3000-KILOWATT  TURBINE  (1500  R.P.M.). 


Brake 

Steam  Pressure 

Superheat, 

Vacuum, 

Pounds  per 

Horsepower. 

(Gauge). 

Deg.  F. 

Inches. 

B.H.P.  Hour. 

2295 

152.0 

102 

26.2 

12.36 

4410 

143-9 

87 

26.2 

11.85 

Rateau  Turbine.     A  looo-kilowatt  Rateau  turbine  built  at  the 
Oerlikon  works  gave  the  following  results  of  steam  consumption : 


Kilowatts. 

Steam  Pressure 
(Absolute). 

Superheat, 
Deg.  F. 

Vacuum, 
Inches. 

Pounds  per 
Kilowatt-hour. 

194 

425 
87I 

1024 

186 

155 
181 
179 

47 

21 
II 

10 

27.73 
27.6 
23.6 

25-05 

31.97 
24.91 
24.69 
21.98 

270 


THE   STEAM   TURBINE 


Zoelly  Turbine.  A  55oo-kilowatt  Zoelly  turbine  installed  at 
the  Quest  Electricity  Works,  Paris,  is  said  to  operate  at  full  load 
with  a  steam  consumption  of  approximately  12.0  pounds  per 
brake  horsepower-hour  at  160  pounds  per  square  inch  gauge 
pressure,  200  degrees  F.  superheat,  and  27  inches  vacuum. 

De  Laval  Turbine.  The  following  table  gives  results  of  tests 
by  Dean  &  Main  of  a  3oo-horsepower  De  Laval  turbine: 


Brake 
Horsepower. 

Steam  Pressure 
(Gauge). 

Superheat. 
Deg.  F. 

Vacuum, 
Inches. 

Pounds  per 
B.H.P.  Hour. 

196.0 
298.4 
352.0 

197.7 
197.0 
198.5 

16 
64 
84 

27.4 
27.4 
27.2 

15.62 
14-35 
13-94 

The  results  shown  in  the  above  tests  give  the  relative  steam 
economy  of  the  principal  types  of  turbines  from  light  load  to 
overload.  Tables  I  and  II  *  on  the  following  page  give  the  com- 
parative results  of  the  latest  reported  tests  in  America  and  in 
Europe. 

HEAT    UNIT    BASIS    OF   EFFICIENCY. 

The  usual  methods  used  for  correcting  steam  turbine  tests  to 
get  a  standard  for  comparison  explained  in  Chapter  VI  are  not 
established  on  a  highly  scientific  basis.  Engineers  appreciate 
generally  that  a  more  rational  method  of  comparison  of  the 
economy  of  heat  engines  on  a  heat  unit  basis  should  be  adopted 
in  cases  where  it  is  practicable.  As  regards  steam  turbines  there 
are,  however,  so  many  uncertain  factors  entering  into  the  deter- 
mination of  a  thermodynamic  efficiency  from  the  available  energy 
that  for  the  present  such  methods  can  be  of  little  value,  except 
in  some  special  cases.  Comparatively  high  superheats  are  now 
generally  used,  and  our  knowledge  of  the  effect  of  reheating  in  a 
multi-stage  turbine  is  very  indefinite. 

A  thermal  efficiency  can,  however,  be  calculated  readily  and 

more  satisfactorily  by  determining  what   percentage  the  heat 

equivalent  of  the  work  is  of  the  heat  "used  by  the  turbine/' 

assumed  to  be  the  difference  between  the  total  heat  in  the  steam 

*  Compiled  by  H.  T.  Herr  and  A.  G.  Christie. 


TESTS   OF   STEAM   TURBINES 


271 


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272  THE   STEAM   TURBINE 

at  the  initial  conditions  and  the  heat  ("of  the  liquid")  in  the 
condensed  steam  at  the  temperature  of  the  exhaust. 

By  this  method  the  full  load  test  of  a  Westinghouse-Parsons 
turbine  reported  by  F.  P.  Sheldon  &  Co.  will  be  calculated  from 
the  data  given  in  an  official  report. 

In  order  to  make  the  results  of  such  calculations  of  steam 
turbine  tests  comparable  with  the  usual  heat  unit  computations 
of  reciprocating  steam  engine  tests  the  results  are  generally 
expressed  in  terms  of  indicated  or  "  internal "  horsepower. 
F.  P.  Sheldon  &  Co.  assumed  the  mechanical  efficiency  of  a 
reciprocating  engine  of  about  the  same  capacity  at  full  load  to 
be  93.3  per  cent. 

THERMAL  EFFICIENCY  OF  A  400-KILOWATT  STEAM  TURBINE. 

Brake  horsepower 660 

Corresponding  indicated  or  "internal"  horsepower  of  a  recip- 

660 
rocatmg  engine   =  708 

'^OO 

Total  steam  used  per  hour,  pounds 9169 

Steam  used  per  "  internal"  horsepower  per  hour,  pounds 12.96 

Steam  pressure,  pounds  per  square  inch  absolute 166.9 

Superheat,  degrees  F 2.9 

Vacuum,  referred  to  30  inches  barometer,  inches 28.04 

Temperature  of  condensed  steam,  degrees  F.  (at  ,96  pound  per 

square  inch  absolute  pressure) 100. 6 

Total  heat  contents  of  one  pound  of  dry  saturated   steam  at 

the  initial  pressure,  B.T.U 1193 . 9 

Heat  equivalent  of  superheat  in  one  pound  of  steam,  B.T.U. 

(Cp.  from  Fig.  30) 1.9 

Total  heat  contents  of  one  pound  of  superheated  steam,  B.T.U.  1195.8 

Heat  of  liquid  in  condensed  steam,  B.T.U 68.6 

Heat  used  in  turbine  per  pound  steam,  B.T.U 1127.2 

Heat  used   in  turbine  per  "internal"  horsepower   per 

minute,  B.T.U.  =  1127.2  X  -^-  =     243.5 

Heat  equivalent  of  one  horsepower  per  minute,  B.T.U.  =  3  42 .42 

770 

Thermal  efficiency,  per  cent.  (42.42  -r-  243.5) *7  4 

Standard  forms  for  data  sheets  and  for  tabulating  results  of 
steam  turbine  tests  are  given  in  Power  Plant  Testing  by  the 
author.  (See  pages  315-340.)  Full  explanations  of  methods 
and  of  necessary  precautions  are  given. 


CHAPTER   XIV. 
STEAM  TURBINE   ECONOMICS. 

The  Best  Conditions  of  Vacuum,  Superheat,  and  Steam  Pressure. 
For  normal  operating  conditions,  a  great  deal  can  be  learned 
about  the  most  profitable  and  satisfactory  vacuum,  superheat, 
and  initial  steam  pressure  for  steam  turbines  from  a  comparison 
and  study  of  existing  modern  power  plants. 

For  this  purpose  a  table  *  is  given  on  the  following  page  in 
which  data  are  given  regarding  the  vacuum,  superheat,  and  steam 
pressure  of  a  large  number  of  steam  power  plants.  This  table 
is  compiled  from  fifty-eight  turbine  plants  in  America  and  in 
England.  The  figures  represent  the  number  of  plants  working 
under  the  conditions  stated  at  the  head  of  each  column. 

There  is  no  doubt  that  such  comparative  data  of  operating 
conditions  are,  from  a  practical  viewpoint,  of  considerable  impor- 
tance. Although  these  figures  were  collected  in  1904  and  1905, 
they  may  be  taken  to  represent  very  well  the  average  practice  of 
the  last  few  years  as  well,  except  that  in  America  there  has  been 
a  tendency  to  operate  a  larger  percentage  of  the  plants  with 
from  100  to  150  degrees  F.  superheat  and  at  about  28  inches 
vacuum. 

The  Question  of  the  Most  Profitable  Vacuum.  Steam  turbine 
manufacturers  are  inclined,  naturally,  because  of  the  obvious 
advantages  of  turbines  over  reciprocating  engines  for  operation 
at  high  vacuums,  to  draw  attention  to  the  reduction  in  the  steam 
consumption  when  a  plant  is  operated  at  a  high  vacuum.  Then 
the  question  is  often  raised  as  to  the  actual  economy  considering 
the  increased  first  cost  of  the  condensers,  pumps,  and  piping, 

*  J.  R.  Bibbins,  in  the  Report  of  American  Street  Railway  Association,  October, 
1904,  page  201,  and  from  data  collected  in  1905  by  Messrs.  Stevens  and  Hobart. 

273 


2/4 


THE  STEAM  TURBINE 


together  with  probably  larger  operating  expenses.  A  great  deal 
depends  on  the  local  conditions,  particularly  on  the  average 
temperature  of  the  condenser  cooling  water.  At  places  only 
slightly  elevated  above  the  sea-level  and  where  the  temperature 
of  the  water  supply  for  the  condensers  is  very  low  —  near  the 


Limits  of 
Capacity  in 
Rated  Kilo- 
watts of  Plant. 

Character  of 
Service. 

Limits  of 
Vacuum, 
Inches  of 
Mercury. 

Limits  of 
Superheat, 
Degrees  F. 

Limits  of 
Steam  Press., 
Pounds  Gauge. 

00 
IN 

§ 

2 

oo 

(N 

2 

<M 

0 

o 

0 

o' 

2 
o 

0 

6 

0 

2 

0 

d 

0 

I 

0 

o 

2 

0 

o 

d 

2 

10 

2 

o 

40,000  
25,000  to  10,000 

10,000  to  5,000. 
5,000  to  3,000.. 

4,OOO  tO   2,000.. 

2,000  to  i  ,000.  . 
Below  1,000.  .  . 
Totals 

Traction  .  .  . 
(Traction..  .  . 
•{Light     and 
1     power.  .  .  . 
Traction..  .  . 
Light     and 
power.  .  .  . 
(Traction.  .  . 
Power 

I 

2 

I 

I 

4 

2 
2 

I 

2 

2 
j 

i 
i 

2 

4 

i 

i 

2 

I 

2 

I 

I 
4 

2 
2 

I 

2 
I 

2 

... 

2 

2 

I 
2 

2 

Light     and 
power 

3 

{Traction  . 

I 

2 

i 

2 

2 

I 

Power  
Light     and 
power 

4 
4 

4 

4 

Traction 

1  Power 

14 

14 

14 

4 

1  Light     and 
I     power 

4 

4 

5 

37 

II 

5 

2 

14 

12 

10 

20 

IO 

9 

3° 

9 

freezing  point  for  a  large  part  of  the  year  —  it  is  doubtless  profit- 
able to  install  condensing  apparatus  of  sufficient  size  to  operate 
steam  turbines  at  from  28.5  to  29  inches  vacuum.  The  following 
table,  calculated  by  J.  R.  Bibbins,  gives  side  by  side  the 
theoretical  and  the  practical  vacuums  at  sea-level  for  varying 
temperatures  of  the  cooling  water. 


STEAM  TURBINE  ECONOMICS 


275 


VACUUM  AT  SEA  LEVEL  FOR  VARYING  TEMPERATURES  OF  COOLING  WATER 


Temperatures  of  Cooling 
Water.      Deg.  F. 

Theoretical 
Possible 
Vacuum. 
Inches. 

Perfect  Con- 
denser, No 
Temperature 
Difference. 
Inches. 

Actual  Con- 
denser,  15°  F. 
Difference. 

Inches. 

Actual  Con- 
denser,  15°  F. 
Difference. 
Inches. 

Ratio  Water  to  Steam  .... 

Infinite. 

60  to  i. 

60  to  i. 

100  to   i. 

32 
60 
70 

75 

29.83 
29.50 
29.30 
29.  10 

29.67 

29.  12 
28.77 
28.51 

29-43 
28.56 
27.72 
27-37 

29-54 
28.82 
28.38 
28.11 

In  modern  surface  condenser  installations  there  is  usually  a 
difference  of  about  15  degrees  F.  between  the  temperature  of  the 
condensed  steam  and  of  the  discharged  water.  It  will  be  seen 
then  in  the  above  table  that  with  the  reasonable  ratio  of  cooling 
water  to  steam  of  60  to  i  the  maximum  vacuum  obtainable,  when 
the  cooling  water  is  taken  in  at  60  degrees,  is  28.6  inches,  and 
when  taken  in  at  70  degrees  is  only  27.7  inches.* 

The  fact  must  not  be  lost  sight  of  that  the  elevation  has  an. 
appreciable  effect  on  the  maximum  possible  vacuum  and  con- 
sequently on  the  most  profitable  vacuum.  At  an  elevation  of 
1000  feet  above  the  sea-level  the  possible  vacuum  obtainable 
with  a  given  condensing  apparatus  will  be  about  an  inch  less  than 
at  tide-water,  and  the  vacuum  reduction  is,  of  course,  in  pro- 
portion for  other  elevations. 

Bibbins  has  also  calculated  the  actual  percentage  saving  when 
the  condenser  equipment  is  increased  so  that  the  plant  can  be 
operated  at  28  inches  instead  of  26  inches.  It  is  estimated  that 
the  cost  of  the  condenser  equipment  including  pumps  and  piping 
will  be  $4000  more  for  a  2ooo-kilowatt  plant  to  operate  at  28 
inches  vacuum  than  at  26  inches  vacuum.  The  results  are  given 
in  the  table  on  the  following  page : 

*  A  firm  of  engineers  which  has  been  installing  steam  turbines  almost  exclusively 
in  the  power  plants  it  has  designed  and  constructed,  has  equipped  a  power  plant  at 
Tampa,  Florida,  with  Diesel  oil  engines  because  of  the  cost  of  cooling  water  in  a 
warm  climate. 


2/6 


THE  STEAM  TURBINE 


RELATIVE  ECONOMY  OF  28  INCHES  VACUUM  OVER  26  INCHES  IN  A 
2000-KILOWATT  PLANT. 

Estimated  Increased  Cost  of  Equipment  is  $4,000. 


Water 

Net  Saving 
expressed  as 
Percentage  of 
Increased  Cap- 
ital Cost  to 
Secure  28  Ins. 
Vacuum  over 
that  for  26  Ins. 

Average 
Load 
in 
Kilowatts. 

Hours  of 
Service 
per 
Day. 

Actual 
Evapo- 
ration, 
Pounds. 

Steam 
Consump- 
tion, 
Average 
Pounds 
per  Kilo- 
watt-hour. 

Saved  per 
Kilowatt- 
hour 
by  Rais- 
ing Vac- 
uum from 
26  Ins. 

Coal, 
Dollars 
per 
Ton. 

to  28  Ins. 

118 

1500 

24 

9-5 

23 

1.84 

4-5° 

27 

1000 

24 

8 

22 

l.76 

2.25 

4 

1000 

10 

8 

22 

1.76 

I-I3 

In  the  calculations  for  the  above  results  the  rate  of  interest  was 
taken  at  5  per  cent,  and  depreciation  at  7.5  per  cent,  on  the  extra 
cost  of  equipment.  Cost  of  extra  power  consumed  was  at  the 
rate  of  i  cent  per  kilowatt-hour,  and  10  cents  per  1000  gallons  of 
feed- water  saved. 

Although  it  may  be  stated,  in  general,  that  it  is  profitable  to 
equip  a  station  to  operate  under  normal  conditions  at  a  vacuum 
of  28  inches  instead  of  26  inches,  it  will  be  observed  from  the 
above  table  that  there  are  cases  where  there  is  practically  no 
advantage  either  way.  In  the  third  case  given,  where  the  plant 
has  only  a  lo-hour  load  and  coal  is  cheap,  the  gain  is  only  4  per 
cent. 

Operation  at  29  inches  vacuum  compared  with  28  inches  is 
not  nearly  so  favorable  to  the  higher  vacuum  as  the  comparison 
of  28  inches  with  26  inches. 

It  will  be  observed  in  the  table  on  the  following  page  that  the 
volume  of  the  steam  is  increased  practically  in  the  same  ratio  (the 
volume  is  practically  doubled)  when  the  vacuum  is  increased  from 
28  inches  to  29  inches  as  when  increased  from  26  inches  to  28 
inches.  Fig.  181  shows  graphically  the  very  large  increase  in 
volume  of  the  steam  in  its  passage  through  the  five  stages  of  a 
large  Curtis  turbine  operating  at  29  inches  vacuum. 


STEAM  TURBINE  ECONOMICS 


277 


TABLE  OP  THE  VOLUME  OF  A  POUND   (SPECIFIC  VOLUME) 
OF    DRY  SATURATED   STEAM   AT    HIGH   VACUUMS. 


Vacuum,  Inches. 

Volume,  Cubic  Feet. 

29 
28 
26 

665* 
342 
I76 

*  Ratio  of  the  volume  at  28  inches  vacuum  to  that  at  26  inches  is  i  .94,  and  the  ratio  of  volumes 
at  29  inches  vacuum  and  at. 28  inches  is  i  .95 . 

It  may  be  stated  then  that  the  capacity  of  the  condensing 
equipment  for  a  turbine  operating  at  29  inches  vacuum  must  be 
practically  four  times  as  large  as  it  would  be  for  one  exhausting 
at  about  26  inches  vacuum.  Or,  in  other  words,  the  volume  of 
a  pound  of  steam  at  the  exhaust  is  166  cubic  feet  larger  at  28 
inches  vacuum  than  at  26  inches,  but  that  at  29  inches  it  is  323 
cubic  feet  larger  than  at  28  inches.  Now  if  as  has  been  stated 
the  cost  of  the  condensing  equipment  is  $4000  more  for  a  2000- 
kilowatt  unit  when  28  inches  vacuum  is  substituted  for  26  inches, 
the  increased  cost  is  obviously  much  greater  when  29  inches 
vacuum  is  compared  with  28  inches. 

For  turbines  of  which  the  steam  consumption  is  not  reduced 
very  much  more  per  inch  of  vacuum  between  28  and  29  inches 
than  between  26  and  28  inches,  in  a  comparison  of  the  economic 
operation  at  29  inches  vacuum  with  28  inches  there  is  a  large 
increased  capital  cost  for  condensing  equipment  which  is  not 
offset  by  a  proportionate  reduction  of  the  steam  consumption, 
and  there  are  probably  comparatively  few  places  —  unusually 
located  as  regards  low  elevation,  low  temperatures,  large  capacity, 
expensive  fuel,  or  high  load  factor  —  where  an  installation  for 
operation  at  an  average  vacuum  of  29  inches  is  profitable. 

The  percentage  change  in  the  steam  consumption  is  approxi- 
mately the  same  at  light  ("fractional")  loads  as  at  full  load  (see 
page  130).  Now  because  the  steam  consumption  per  kilowatt- 
hour  is  greater  at  light  loads,  the  change  in  steam  consumption 
per  kilowatt-hour  is  therefore  also  greater  at  light  loads  than  at 


278 


THE  STEAM  TURBINE 


full  load.  Ordinarily  this  fact  is  stated  by  saying  that  a  change 
in  vacuum  has  a  greater  "effect"  at  light  than  at  full  load,  and 
that  the  effect  is  more  marked  at  high  than  at  low  vacuums. 
Effect  of  vacuum  on  the  steam  consumption  of  any  impulse 


Percentage  Decrease  lu  Steam  Consumption 
per  Inch  Increase  in  Vacuum 
f*  r  J®  i»  S*  P5  .*• 

OCHOCT  OCTOCTO 

/ 

/ 

f 

/ 

/ 

/ 

/ 

/ 

A 

/ 

/ 

/ 

/ 

. 

/ 

x 

/ 

X 

x 

^ 

x 

^> 

^ 

) 

.—  = 

! 

= 

< 

—  — 

••  ' 

t 

—  • 

—  —  ! 

^^ 

~ 

^ 

10 

^^* 

~1 

2 

~ 

T~ 

~ 

~ 

~T 

r~ 

~ 

~ 

~ 

2~ 

~ 

~ 

~ 

T" 

~ 

~ 

- 

Vacuum  in  Inches  of  Mercury 

FlG.  190.     Percentage  Curve  of  the  Effect  of  Vacuum  on  the  Steam  Consumption 
of  a  Single-Stage  Impulse  Turbine. 


turbines  of  the  single-stage  type  is  probably  shown  very  accurately 
by  Fig.  190  (reproduced  from  Fig.  88).* 

*  In  the  catalogs  of  the  General  Electric  Company  it  is  stated  that  a  curve  like 
Fig.  127  is  typical  for  most  Curtis  turbines.  Actually,  however,  a  curve  like 
Fig.  78  is  more  accurate.  Emmett  has  stated  recently  in  a  published  communi- 
cation that  "around  27  inches  the  change  in  economy  per  inch  is  6.6  per  cent.; 
28  inches  7.8  per  cent.;  and  29  inches  9.5  per  cent." 

Parsons  states  in  a  paper  read  before  the  Institution  of  Electrical  Engineers  in 
1904  that  in  a  turbine  "the  benefit  derived  from  a  good  vacuum  is  much  more  than 
in  a  reciprocating  engine.  Every  inch  of  vacuum  between  23  and  28  inches  affects 
the  steam  consumption  on  an  average  about  3  per  cent,  in  a  loo-kilowatt;  4  per  cent, 
in  a  5oo-kilowatt;  and  5  per  cent,  in  a  i5oo-kilowatt  turbine,  the  effect  being  more  at 
high  vacuum  and  less  at  low."  It  seems  very  doubtful  to  the  author  whether,  in 
general,  vacuum  corrections  can  be  classified  according  to  the  size  of  the  turbine. 
There  are  some  very  large  turbines  of  the  Parsons  type  of  which  the  vacuum 
correction  is  less  than  4  per  cent,  per  inch  of  vacuum. 


STEAM  TURBINE  ECONOMICS 


279 


The  variation  of  the  steam  consumption  of  a   5oo-kilowatt 
Westinghouse-Parsons  turbine  for  vacuums  from  25  to  29  inches 


FIG.  191.     Curves  of  Steam  Consumption  of  a  5oo-Kilowatt  Westinghouse-Parsons 
Turbine  for  25,  26,  27,  28,  and  29  Inches  Vacuum. 


from  light  loads  to  overloads  is  illustrated  by  the  curves  in  Fig.  191. 
What  might  be  called  a  curve  of    normal  vacuum  correction 


280 


THE    STEAM   TURBINE 


factors  for  comparing  those  of  26,  27  and  29  inches  with  28  inches 
in  Westinghouse  turbines  is  given  in  Fig.  192. 

Chilton,*  after  stating  that  the  impression  is  no  longer  so 
common  that  a  high  vacuum  is  necessary  to  secure  good  results 
with  steam  turbines,  says  that  the  difference  in  economy  of  Allis- 
Chalmers-Parsons  turbines  between  24  and  27  inches  vacuum 
is  5  per  cent,  per  inch.  Between  27  and  28  inches  the  saving  is 
6  per  cent.,  and  between  28  and  29  inches  is  7  per  cent. 


300> 


ill 


FIG.  192.     Vacuum  Correction  Factors  for  Westinghouse  Single-Flow  Turbines. 

An  idea  of  the  relative  quantity  of  condensing  water  required 
for  different  vacuums  may  be  gained  by  comparing  that  required 
for  the  usual  operating  vacuums.  For  example,  with  injection 
water  of  70  degrees  F.,  the  usual  temperature  upon  which  con- 
denser guarantees  are  based,  it  is  customary  to  estimate  that 
to  obtain  a  vacuum  of  27  inches  about  36  pounds  of  water  will  be 
used  for  each  pound  of  steam  condensed,  and  about  1.4  times  this 
quantity  is  required  for  a  vacuum  of  28  inches.  With  injection 
water  at  60  degrees  F.,  which  may  be  considered  the  winter  tem- 
perature, the  quantities  required  for  the  foregoing  vacuums  are 

*  Street  Railway  Journal,  Oct.  19,  1907. 


STEAM  TURBINE  ECONOMICS 


28l 


approximately  28  and  34  pounds  respectively.  Having  the  quan- 
tity of  condensing  water  required,  the  cost  of  fuel,  and  cost  of 
water  delivered  to  the  condenser,  the  vacuum  best  suited  to  the 
conditions  under  consideration  may  be  readily  determined.  Theo- 
retically, the  effect  upon  the  turbine  of  reducing  the  vacuum 
below  that  for  which  it  is  designed,  is  to  reduce  the  capacity  and 
to  lower  the  rating  at  which  maximum  economy  is  obtained. 

The  following  table  *  illustrates  the  percentage  gain  in  economy 
per  inch  of  vacuum  for  various  vacuums.  The  close  agreement 
between  the  actual  results  and  the  theoretical  values  should  be 
observed.  The  table  applies,  however,  only  to  turbines  using 
very  high  steam  pressures  and  superheats.  For  "  land  "  turbine 
practice  it  is  serviceable  only  for  comparison. 


Inches  of  Vacuum. 

Gain  in  Per  Cent. 

28 

27 

26 

25 

Curtis  1  

5-i 

5-° 
3-J4 
5-2 

4.8 
4.0 
3-05 
4.4 

4.6 

3-5 

2.Q5 

3-7 

4.2 
3-o 
2.87 
3-0 

Parsons 

\Vestinghouse-Parsons 

Theoretical 

Effect  of  Superheating  on  Economy.  The  effect  of  superheat 
on  the  economy  of  De  Laval  and  Parsons  turbines  is  usually 
stated  to  be  10  per  cent,  per  100  degrees  F.  superheat.  This 
statement  is  probably  very  nearly  correct  for  the  usual  ranges  of 
superheat  in  practice  and  is  the  usual*  correction  employed  by 
most  consulting  engineers  for  correcting  steam  turbine  tests  f  up 
to  about  150  degrees  F.  superheat. 

Some  investigations  made  by  Professor  Hobart  show  that  the 
mean  superheat  correction  for  Parsons  turbines  is  almost  exactly 

*  Mechanical  Engineer,  Feb.  24,  1906. 

f  The  superheat  corrections  used  by  the  engineers  of  the  Westinghouse  Com- 
panies, by  Dean  &  Main,  and  by  Parsons,  are  all  approximately  10  per  cent,  per  100 
degrees  F.  superheat. 


282  THE  STEAM  TURBINE 

10  per  cent,  per  100  degrees  superheat  for  all  superheats  from 
o  to  100  degrees  F.  Between  100  and  150  degrees  superheat  it 
is  approximately  8  per  cent.,  and  between  150  and  250  degrees 
is  about  6  per  cent.  It  is  the  opinion  of  the  author  that  the 
results  of  this  investigation  can  be  considered  quite  accurate,  as 
a  large  number  of  tests  were  compared.  A  curve  showing 
approximately  the  same  sort  of  variation  in  the  superheat  correc- 
tion of  De  Laval  turbines  is  given  in  Fig.  87.  Chilton  states  that 
tests  of  Allis-Chalmers-Parsons  turbines  show  that  the  "incre- 
ment of  saving  becomes  smaller  as  the  superheat  is  increased"; 
adding  that  for  50  degrees  F.  superheat  the  steam  consumption 
is  reduced  7  per  cent,  (at  the  rate  of  14  per  cent,  per  100  degrees) ; 
for  100  degrees  10  per  cent;  and  for  150  degrees  12.5  per  cent, 
(at  the  rate  of  a  little  more  than  8  per  cent,  per  100  degrees). 

According  to  Kruesi  of  the  General  Electric  Company  IOG 
degrees  F.  superheat  reduces  the  steam  consumption  of  Curtis 
turbines  8  per  cent.,  but  "the  first  50  degrees  of  superheat  is  of 
greater  value  than  the  second  50  degrees."* 

When  steam  at  about  150  pounds  per  square  inch  gauge 
pressure  is  superheated  100  degrees  F.  the  total  heat  of  the 
steam  is  increased  about  4.8  per  cent,  with  an  additional  fuel 
expenditure  of  approximately  6  per  cent,  if  the  boiler  equipment 
is  good.  Now  since  the  steam  consumption  is  reduced  from  8  to 
10  per  cent,  for  100  degrees  F.  superheat  there  is  obviously  a 
saving  of  from  2  to  4  per  cent,  in  the  cost  of  fuel. 

Experience  seems  to  show  that  the  best  economic  results  will 
be  obtained  with  from  100  to  150  degrees  F.  superheat  for  tur- 
bines of  the  Parsons  type,  and  about  50  degrees  superheat  for 
Curtis  turbines  of  more  than  one  stage.  In  all  kinds  of  turbines 
of  the  single-stage  impulse  type  there  is  probably  always  a  saving 

*  Because  curves  of  steam  consumption  per  kilowatt-hour  for  varying  super- 
heats (like  Fig.  126)  were  apparently  straight  lines,  most  turbine  engineers,  until 
very  recently,  believed  that  at  high  superheats  the  percentage  correction  was 
increased  instead  of  being  reduced  as  more  recent  results  show.  Since  it  has 
been  fairly  well  established  that  the, specific  heat  of  superheated  steam  has  very  low 
and  minimum  values  at  from  about  200  to  250  degrees  F.  superheat,  the  later 
results  seem  to  be  the  more  reasonable. 


STEAM  TURBINE  ECONOMICS  283 

of  from  4  to  5  per  cent,  in  fuel  cost  per  100  degrees  F.  superheat 
within  the  practicable  limits  of  superheating. 

Although  there  is  much  yet  to  be  determined  concerning 
superheated  steam,  it  has  been  shown  by  experience  in  turbine 
plants  that  a  considerable  saving  in  fuel  can  be  secured  by  super- 
heating the  steam  at  least  a  moderate  amount.  The  greater 
saving  in  turbines  of  the  Parsons  type  over  multi-stage  Curtis 
turbines  is  due  to  the  larger  "skin-friction"  or  disk  and  blade 
rotation  losses  of  the  large  number  of  rows  of  blades  in  Parsons 
turbines.  The  curves  in  Fig.  69  show  the  very  large  percentage 
that  these  losses  are  reduced  when  the  blades  revolve  in  dry 
steam  instead  of  wet  steam.  When  the  admission  steam  to  a 
Parsons  turbine  is  dry  saturated  the  steam  in  the  low-pressure 
stages  will  probably  have  nearly  20  per  cent,  of  moisture,  while 
if  it  is  superheated  1 50  to  200  degrees  F.  the  steam  in  these  stages 
will  be  nearly  dry. 

Finally,  the  use  of  a  high  degree  of  superheat  must  depend  not 
only  on  the  type  of  turbine,  the  load  factor,  and  the  size  of  the 
units  but  also  upon  the  nature  of  the  service  as  regards  severe 
and  frequent  variations  in  the  load,  having  in  mind  the  difficulties 
which  have  been  encountered  in  the  practical  operation  of  super- 
heaters, steam  piping,  valves,  pumps,  and  auxiliary  machinery. 

Reasons  for  the  Improved  Economy  in  Turbines  and  Recipro- 
cating Engines  Due  to  Superheated  Steam.  A  gain  in  steam  and 
fuel  economy  results  from  the  use  of  superheated  steam  in  either 
turbines  or  reciprocating  engines.  In  the  turbine  the  gain  comes 
principally  from  the  reduced  fluid  friction  of  the  steam  moving 
at  a  high  velocity  through  passages  and  blades,  some  of  which 
have  also  a  comparatively  high  velocity.  In  a  reciprocating 
engine  the  gain  from  superheated  steam  is  due  to  the  reduction  of 
cylinder  condensation,  resulting  in  less  loss  due  to  the  cooling  of 
the  cylinder  from  the  reevaporation  of  moisture  at  the  lower 
pressures  near  the  end  of  the  stroke.  On  account  of  this  cooling 
of  the  cylinder  ends,  the  loss  due  to  the  "initial  condensation" 
of  the  steam  admitted  on  the  return  stroke  is  often  40  to  50  per 
cent,  of  the  weight  of  steam  admitted.  This  loss  is  partly  or 


284  THE  STEAM  TURBINE 

entirely  prevented  when  the  steam  is  superheated,  depending 
upon  the  degree  of  superheat.  In  a  steam  turbine  there  is  a 
similar  loss  due  to  condensation,  but  it  is  due  almost  entirely  to 
the  mere  expansion  of  the  steam.  The  walls  of  the  turbine 
casing  remain,  however,  at  a  practically  uniform  temperature,  so 
that  there  is  no  opportunity  for  loss  through  reevaporation  of 
condensed  steam. 

Steam  Pressure  Best  Suited  to  Turbines.  It  is  the  general 
opinion  of  practical  engineers  that  probably  the  most  economical 
operating  pressure  for  the  usual  power-house  services  is  about 
150  pounds  per  square  inch  gauge  pressure  (165  absolute)  at  the 
throttle  valve,  and  that  a  greater  saving  can  always  be  obtained 
by  the  use  of  a  moderate  amount  of  superheat  than  by  increasing 
the  pressure  beyond  this  point. 

Chilton  states  that  there  is  a  gain  of  2  per  cent,  in  steam  con- 
sumption from  increasing  the  steam  pressure  from  150  to  175 
pounds  per  square  inch  *'  and  i  per  cent,  for  an  increase  from 
175  to  200  pounds  per  square  inch.  But  against  the  saving  in 
fuel  due  to  a  reduced  steam  consumption  must  be  charged  the 
increased  cost  of  piping,  valves,  and  boilers,  and  also  the  loss  due 
to  increased  leakage.  Increasing  the  steam  pressure  will  also 
increase  considerably  the  cost  of  the  turbine.  A  "rough  and 
ready"  correction  used  a  great  deal  by  turbine  engineers  is  one- 
tenth  per  cent,  per  pound. 

Speed  Variation  as  it  Affects  Economy.  A  steam  turbine  will 
give  its  best  economy  at  some  particular  speed,  just  as  it  has  been 
found  to  give  its  best  economy  at  some  definite  load.  For  this 
reason  the  design  of  a  turbine  should  be  worked  out  very  carefully 
with  velocity  diagrams  to  determine  whether  at  the  speed  required 
by  the  operating  conditions  it  will  give  the  best  economy.  When- 
ever any  changes  are  made  in  the  design  of  a  turbine,  the 
manufacturers  will  always  make  tests  to  determine  the  steam  con- 

*  The  engineers  of  the  Westinghouse  and  General  Electric  companies  use 
practically  the  same  correction  for  initial  pressure.  It  may  be  added  that  the 
correction  for  exhaust  pressure  (back  pressure)  of  non-condensing  turbines  is  about 
ten  times  as  large  as  the  correction  for  initial  pressure. 


STEAM  TURBINE  ECONOMICS  285 

sumption  at  various  speeds,  and  curves  like  those  shown  in  Fig.  80 
are  calculated  and  plotted.  If  it  is  found  that  the  turbine  has  a 
lower  steam  consumption  at  a  slightly  different  speed  from  that 
for  which  it  is  rated,  either  the  angles  of  the  blades  or  the  pressures 
must  be  changed.  The  reasons  for  such  changes  are  obvious, 
because  the  blade  speed  has  a  very  definite  relation  to  velocity 
of  the  steam  in  the  blades.  If  the  designer  is  not  successful  in 
securing  this  relation  for  the  rated  speed,  there  will  be  impact  of 
the  steam  against  the  blades  and  a  consequent  loss  of  efficiency. 

The  curve  of  steam  consumption  in  Fig.  80  shows  the  change 
in  economy  at  various  speeds.  At  2000  revolutions  per  minute 
the  steam  consumption  is  19.6  pounds  per  kilowatt-hour;  at 
1800  revolutions  (rated  speed)  it  is  19.45  pounds;  at  1600 
revolutions,  about  19.8  pounds;  at  1400  revolutions,  about  20.7 
pounds;  and  at  1000  revolutions,  about  24.7  pounds.  It  will  be 
observed  in  these  curves  that  the  ideal  conditions  have  been 
secured  in  the  design  of  this  turbine;  that  is,  the  steam  consump- 
tion is  lowest  and  the  output  (load)  greatest  at  the  rated  speed. 
Within  a  range  of  about  50  revolutions  above  or  below  the  rating 
(a  total  variation  of  about  6  per  cent.)  the  steam  consumption  is 
practically  constant.  These  curves  are  typical  for  all  good 
designs  of  steam  turbines. 

When  a  speed  test  is. made  of  an  impulse  turbine  the  best 
results  are  obtained  as  regards  the  accuracy  of  the  design  by 
running  the  turbine  with  a  number  of  nozzles  wide  open  to  give 
approximately  full  load.  The  test  for  each  speed  can  then  be 
made  of  comparatively  short  duration,  as  the  steam  can  be  weighed 
continuously  between  the  first  and  last  tests  without  interruption 
when  the  speed  is  being  changed.  With  a  constant  number  of 
nozzles  discharging  steam  the  rate  of  flow  will  be  the  same  at  all 
speeds. 

Comparative  Economy  of  Steam  Turbines  and  Reciprocating 
Engines.  To  summarize  the  results  of  tests  on  a  number  of 
large  steam  turbines  and  reciprocating  engines  the  following 
tables  have  been  prepared.  Steam  consumption  of  most  of  the 
turbine  tests  was  given  in  the  published  data  in  terms  of  kilowatt- 


286 


THE  STEAM  TURBINE 


hours  or  electrical  horsepower-hours.  In  order  to  make  com- 
parisons with  the  reciprocating  engine  it  was  necessary  to  reduce 
all  to  a  common  standard  —  brake  horsepower-hour.  To  express 
all  the  results  in  this  common  standard  various  efficiencies  must 
be  assumed.  In  the  calculations  the  generator  efficiencies  given 
on  page  362  were  used  to  obtain  the  following  coefficients  to 
change  the  steam  consumptions  from  the  rate  per  kilowatt-hour 
to  that  per  brake  horsepower-hour: 


Rating  of  Turbine, 
Kilowatts. 

Coefficient. 

300  and  400 

500 
1,000  to    3,000 
5,000  to  10,000 

.68 

•71 
.72 

•73 

Mechanical  efficiency  of  reciprocating  engines  of  30x30  to  5000 
horsepower  is  about  91  per  cent.;  1000  horsepower,  about  90  per 
cent. ;  and  400  to  700  horsepower,  about  89  per  cent. 

In  the  following  tables  are  given  the  steam  consumptions  of  a 
large  number  of  steam  turbines  and  some  particularly  good 
reciprocating  engines.  A  great  many  of  the  steam  turbine  tests 
given  are  approximately  the  full  load  data  taken  from  the  tests 
recorded  at  the  end  of  the  preceding  chapter,  and  some  others  are 
taken  from  Chapter  VI. 

The  ratings  given  in  the  tables  are  those  for  what  is  generally 
known  by  engineers  as  "full  load;"  meaning  that  the  turbine 
can  carry  economically  a  load  at  least  50  per  cent,  larger  than 
this  rating.  This  statement  is  necessary  because  some  manu- 
facturers use  a  rating  based  on  maximum  output. 

Assuming  average  values  of  the  corrections  given  above  by 
various  authorities,  an  approximate  equivalent  steam  consumption 
has  been  calculated  for  each  engine  at  o  degrees  F.  superheat, 
28  inches  vacuum,  and  165  pounds  per  square  inch  absolute 
steam  pressure. 


STEAM  TURBINE  ECONOMICS 


287 


STEAM   CONSUMPTION  OF  TURBINES. 

A=  American,   E=  English,   F=  French,   G"=  German,   S=  Swiss,    and  W-P 

Parsons. 


Westinghouse- 


Turbine. 

Rated 
Power. 

Conditions  of  Test. 

Steam  per  Hour 
as  per  Test. 

Equivalent 
Steam  per 
b.hp.-hr.  at 

Super- 
heat, 
Deg.F. 

Vacuum 
Inches. 

Steam 
Pressure 
Lbs. 
Abs. 

r.p.m. 

Pounds 
per  kw. 

Pounds 
per 
b.hp. 

o  Degs. 
Sup.,  28  ins. 
Vac.,    165 
Lbs.  Abs. 
Press* 

15-81 
15-29 

14.71 
13-99 
*3.  77 

15.84 

13.17 

13-20 
14.68 
13.25 
13.04 
12.83 
15-52 
13-48 
13-47 

12.  00 

".95 
".95 

12.00 

De  Laval  (G)  .  .  . 
De  Laval  (G)  .  .  . 
De  Laval  (A).... 

Parsons  (E)  .  .  . 
VV-P(A)  
W-P  (A)  

Zoelly  (G)  

W-P  (A)  
Curtis  (A)  ... 

hp. 
3° 
150 
300 

kw. 
300 
300 
300 

hp. 
500 

kw. 
500 
5°o 
500 
500 

1000 
IOOO 
1200 

1500 

2000 
300O 
5000 
7500 
0000 

0 
0 

o 

0 

5 

JOO 

107 

290 
104 

0 
0 
10 

468 

125 

207 
235 
142 

06 

116 

non-con 
26.4 
26.6 

26.6 
28.0 
28.0 

28.7 

27.8 
28.0 
26.  9 
26.7 
27.0 
25.0 
28.8 
27.5 
28.5 
27.0 
28.8 
27.3 
29.  6 

IOO 

114 
206 

158 

160 

1  68 

201 

164 
165 

168 
136 
163 
179 
178 
144 
181 
139 
189 
i93  . 
194 





39-6 
17.  70 
15-17 

15.  70 
13.99 
12.48 

'3-  37 

13.28 
10.  71 
J4-55 
15-05 
13-61 
15.80 

II.  00 
12.  67 
10.  82 
10.  60 
9.87 
11.03 

9.  ii 

3000 
3600 
3600 

23-  15 

18.82 

15-  10 
20.5 

21.  2 

21.98 
15.30 
17.  60 
15.02 

14.  74 
13-  52 

15.  is 

13.  oo 

3600 
1800 
1800 
2400 
1800 

Curtis  (E) 

Rateau  (F).  .  . 

W-P  (A).  .  . 

Rateau  (S)  
Parsons  (S).  .  .  . 
Parsons  (E)..  .  . 
Curtis  (A)  
Parsons  (G).  .  .  . 
Curtis  (A)  
W-P  (A)  
Curtis  (A) 

000 

1350 
750 
75° 
750 

Additional  data,  from  recent  tests  of  steam  turbines  are  given 
in  tables  I  and  II,  page  271. 

Combined  Steam  Engine  and  Low-pressure  Steam  Turbine. 
The  combination  units  of  a  large  Allis  engine  with  Curtis  ex- 
haust steam  turbines  (see  Fig.  i8ic,  page  25oc),  as  installed  in  the 
59th  Street  Power  Station  of  the  Interborough-Metropolitan 
System  in  New  York,  have  a  rated  capacity  of  15,000  horsepower 

*  Correction  curves  in  Figs.  87  and  88  were  used  to  correct  the  De  Laval  tests  for  superheat  and 
vacuum  and  the  usual  correction  of  .1  per  cent,  improvement  in  economy  per  pound  increase  ot 
pressure. 

For  Parsons  and  Westinghouse-Parsons  turbines  the  following  corrections  were  used: 

Superheat  (300-1000  kw.)  10  per  cent.;  (1200-7500  kw.)  8  per  cent,  per  100  degrees  F. 

Vacuum  (300-1000  kw.)  4  per  cent.;  (1200-7500  kw.)  3  per  cent,  per  inch. 

Pressure  .1  per  cent,  per  pound, 
and  the  following  for  Curtis,  Rateau,  and  Zoelly  turbines: 

Superheat,  8  per  cent,  per  100  degrees  F. 

Vacuum  (26-28  ins.)  7  per  cent.;  (28-29.5  ins.)  8  per  cent,  per  inch. 

Pressure  .1  per  cent,  per  pound. 

t  Referred  to  30  inches  barometer. 


288 


THE   STEAM   TURBINE 


and  give  a  steam  consumption  of  13.19  pounds  per  kilowatt- 
hour  (about  8.74  pounds  per  i.h.p.-hour)  with  steam  supplied 
to  the  engine  initially  dry  saturated  (no  superheat),  194  pounds 
per  square  inch  absolute  pressure  and  exhausting  from  the 
turbine  at  28.8  inches  vacuum,  referred  to  30  inches  barometer. 

STEAM  CONSUMPTION  OF  RECIPROCATING  ENGINES  SHOWING  EXCEPTION- 
ALLY HIGH  ECONOMY. 


Steam  per 

Equiv. 

& 

Hour. 

Steam 

fe 

Consump- 

^ 

•f 

J 

tion  per 

•d 

w 

cq 

ja 

b.h.p.  at 

Engine. 

C 

S 

J 

jj 

I 

r.p.m. 

i.h.p. 

b.h.p. 

oDeg. 
superheat, 

References. 

o3 

• 

$ 

28  Ins. 

• 

- 

£ 

Vac.  and 

R 

§ 

165  Lbs. 

$2 

1 

3 

Abs.  Pres- 

• 

H 

> 

1 

sure. 

Rockwood-Wheelock  .  . 

595 

o 

25-4 

174 

76.4 

13.00 

14.61 

14.62 

F.  W.  Dean,  Trans. 

A.S.M.E.,  1895. 

Mclntosh  &  Seymour. 

1076 

20 

27.1 

138 

99-6 

12.76 

14.19 

13-89 

F.  W.  Dean,  Trans. 

A.S.M.E.,  1898. 

l«avitt  Pumping  En- 

576 

o 

27-3 

191 

Si-6 

11.20 

12-59 

13.03 

E.  F.  Miller,  Tech- 

gine. 

nology  Quarterly, 

Rice  &  Sargent 
(Phila.). 

420 

297 

25.8 

157 

IO2 

9  56 

10.75 

13-39 

D.      'S.       Jacobus, 
Trans.   A.S.M.E., 

1904. 

Westinghouse     (verti- 

5400 

0 

27-3 

200 

76 

11-93 

13-12 

13.76 

Eng.    Record,   May 

cal). 

28,  1904. 

Kerchove  
Mclntosh  &  Seymour 

3600308 
2000!  92 

27.6 
25-5 

146.5 
171 

86 

IOO 

II  78 
11.05 

12.95 

13-58 

Von  der  Kerchove. 

(Boston). 

Allis-Chalmers    (New 

7500 

0 

25-1 

190 

80 

11.96 

York). 

Moabit  (Berlin)  
Erie-Lentz  (simple  en- 

2500 
282 

223 
141 

28.1 

0 

203 
156 

85 

208 

8.96 

15.24* 

il.45t 

J.     A.      Moyer     in 

gine)  . 

Power,  1912. 

Buckeye     (compound 

142 

256 

27.2 

196 

195 

9  65 

I0.82J 

Power,  1912. 

engine). 

1 

Effect  of  Superheat,  Vacuum,  and  Admission  Pressure  on  the 
Economy  of  Reciprocating  Engines.  According  to  Professor 
Schroeter*  the  steam  consumption  of  reciprocating  steam 
engines  is  reduced  about  6  per  cent,  for  50  degrees  F.  and  about 
<)  per  cent,  for  100  degrees  F.  of  superheat.  Parsons  f  has 
shown  that  in  a  triple-expansion  engine  the  steam  consumption 
can  be  reduced  only  .4  per  cent,  per  inch  with  an  increase  of 

*  Storm  Bull.,  Journal  of  Western  Society  of  Engineers,  December,  1903. 
t  Proc.  Insl.  of  Naval  Architects,  April,  1908;  Mechanical  Engineer,  May  i,  1908, 
and  Die  Turbine,  July,  1905. 

J  Probably  World's  records  for  steam  engines. 


STEAM   TURBINE   ECONOMICS  289 

vacuum  between  the  limits  of  25  and  28  inches,  and  at  a  still 
higher  vacuum  there  is  practically  no  gain  at  all.  Increased 
initial  steam  pressure  reduces  the  steam  consumption  of  recip- 
rocating engines  .1  to  .2  per  cent,  per  pound  per  square  inch. 

A  comparison  of  the  two  tables  shows  that  in  large  capacities 
steam  turbines  will  give,  for  the  same  standard  conditions, 
better  economy  than  reciprocating  engines.*  It  is  shown  that 
for  sizes  from  3000  to  9000  kilowatts  the  steam  consumption  of 
turbines  is  about  12  pounds  per  brake  horsepower-hour  at  the 
assumed  standard  conditions  of  o  degrees  superheat,  28  inches 
vacuum,  and  165  pounds  per  square  inch  absolute  pressure;  and 
that,  operating  at  the  same  conditions,  the  steam  consumption 
of  the  best  designs  of  reciprocating  engines  is  about  13  pounds. 

Economy  of  Small  Reciprocating  Engines  and  Turbines. 
Nearly  all  small  high-speed  reciprocating  engines  rapidly  deteri- 
orate in  economy,  primarily  because  the  valve  leakage  becomes 
excessive.  Although  an  engine  of  this  kind  will  meet  the  guar- 
antees of  steam  consumption  in  a  shop  test,  it  has  been  shown 
that  very  soon  they  require  a  much  larger  amount  of  steam.f 
Tests  of  seven  high-speed  engines  of  various  types  rated  at  100 
to  200  horsepower  conducted  by  Dean  and  Wood  in  1907  show 
that  the  steam  consumption  of  such  engines  after  a  comparatively 
short  duration  of  service  was  found  to  vary  from  49.4  to  60.5 
pounds  per  kilowatt-hour  at  full  load.  These  rates  are  very  high 
when  compared  with  the  economy  of  small  De  Laval  and  Curtis 
turbines  as  given  in  Figs.  89  and  128.  Parsons  stated  in  1904 
that  the  full-load  steam  consumption  of  turbine-generators  of  his 
design  under  the  conditions  of  100  degrees  F.  superheat,  27  inches 
vacuum,  and  155  pounds  per  square  inch  absolute  steam  pres- 
sure was  approximately  25  pounds  per  kilowatt-hour  for  one  of 
100  kilowatts  capacity,  while  that  of  the  200  and  500  kilowatt  sizes 

*  It  must  not,  however,  be  overlooked  that  these  standard  conditions  were 
selected  in  the  first  place  for  comparing  the  economy  of  steam  turbines.  It  happens 
that  the  vacuum  is  taken  a  little  higher  than  is  usual  in  the  operation  of  reciprocat- 
ing engines. 

t  "Economy  Tests  of  High  Speed  Engines,"  by  F.  W.  Dean  and  A.  C.  Wood> 
Proc.  American  Soc.  Mech.  Engineers,  June,  1908. 


2QO  THE  STEAM  TURBINE 

was  respectively  22  and  20  pounds.  He  stated  that  the  equiva- 
lent results  with  dry  saturated  steam  and  28  inches  vacuum 
would  be  about  ten  per  cent,  larger.* 

It  has  been  shown  by  repeated  tests  that  the  steam  consump- 
tion of  these  turbines  is  not  materially  increased  when  operated 
continuously  for  long  periods.  Weithammer  f  states  that  he 
made  tests  of  a  De  Laval  turbine-generator  when  new  and  after 
five  years  of  service,  and  calculated  the  deterioration  in  economy 
to  be  not  more  than  two  per  cent. ;  and  this  lower  efficiency  was 
probably  largely  due  to  wear  of  the  reduction  gears.  It  would 
appear  that  the  deterioration  of  Curtis  turbines  should  be  even 
less  because  of  less  erosion  from  steam  at  very  high  velocities 
and  the  absence  of  the  reduction  gears.  It  is  stated  that  there 
are  cases  where  De  Laval  blades  have  been  so  much  worn  as  to 
Tequire  replacing  in  a  year.J  Such  an  experience  is,  however, 
unusual. 

POWER   PLANT   ECONOMICS. 

The  following  table  prepared  by  Mr.  H.  G.  Stott  of  New  York 
is  interesting  in  many  of  its  items.  Actual  data  were  used  to 
determine  the  values  under  the  heads  of  "Maintenance"  and 
"Operation."  The  first  column  is  for  a  plant  with  compound 
condensing  reciprocating  engines  operating  without  superheat, 
and  in  all  cases  the  values  have  been  suitably  corrected  to  make 
the  other  columns  directly  comparable  with  the  first. 

Mr.  Stott  advocates  the  use  of  an  exhaust  steam  turbine  to  be 
•operated  by  the  exhaust  steam  from  reciprocating  engines.  By 
increasing  the  pressure  of  the  steam  supplied  a  moderate  amount 
as  well  as  superheating  it  the  output  of  a  power  plant  of  the  type 
represented  by  the  first  column  in  the  table  can  be  doubled  at  a 
comparatively  small  cost  for  turbines  and  boilers. 

*  Trans.  Inst.  of  Electrical  Engineers,  May,  1904. 

f  Die  Dampfturbinen,  page  104. 

%  Lea  and  Meden,  Transactions  American  Soc.  Mechanical  Engineers,  Vol.  25. 


STEAM  TURBINE  ECONOMICS 


291 


DISTRIBUTION  OF  MAINTENANCE  AND  OPERATION. 
(Charges  per  Kilowatt- Hour.) 


Recipro- 
cating 
Engines. 

Steam 
Turbines. 

Recipro- 
cating 
Engines 
and 
Steam 
Turbines. 

Gas 
Engine 
Plant. 

Gas 
Engines 
and 
Steam 
Turbines. 

Maintenance. 

i.    Engine  room,  mechanical..  .  . 

2-57 

0.51 

1-54 

2-57 

1-54 

2.    Boiler  room  or  producer  room 

4.61 

4.30 

3-25 

I-I5 

i-95 

3.    Coal  and  ash  handling  appa- 

ratus            

0.58 

o.  54 

0.44 

o.  29 

o.  29 

4     Electrical  apparatus        

1  .  12 

I  .  12 

I  .  12 

I.  12 

I  .  12 

Operation. 

5.    Coal  and  ash  handling  labor.  . 

2.26 

2.  II 

1-74 

I-I3 

I-I3 

6     Removal  of  ashes 

I    06 

O   04. 

o  80 

O.  53 

O.53 

7     Dock  rental 

O    74 

O    74 

O.  74 

o.  74 

O.  74 

8     Boiler  room   labor 

7    i? 

6  68 

5.46 

I  •  70 

3.O3 

9.    Boiler  room,  oil,  waste,  etc.  .  . 

0.17 

0.17 

0.17 

0.17 

0.17 

10.    Coal                   

61   30 

IT?      ?Q 

46.87 

26.  31 

25.  77 

n.    Water      

7.  14 

7     TO 

•>-46 

3.  57 

2.  14 

12.    Engine  room,  "mechanical" 

labo'- 

6    71 

I      35 

4  03 

6   71 

4   O3 

1  3     Lubrication 

I    77 

O    35 

I    OI 

I    77 

I.  06 

1  4     Waste   etc 

O    3O 

O    3O 

o  30 

O    3O 

o.  30 

15     "Electric"  labor 

2    52 

2     52 

2    52 

2  .  52 

2  .  52 

Relative  cost  of  maintenance  and 

operation  

IOO    OO 

86  03 

75    72 

^0.67 

46.  32 

Relative  investment  in  per  cent.  .  .  . 

100.00 

82  .  50 

77.00 

100.00 

91.20 

That  the  steam  turbine  plant  has  an  inherent  economy  of  20 
per  cent,  better  than  the  best  type  of  reciprocating  engine  installa- 
tion is  shown  by  a  comparison  of  the  first  and  second  columns. 

Prices  of  Steam  Turbines.  Fig.  206  shows  by  means  of  curves 
the  price  per  kilowatt  of  the  normal  full  load  rating  of  turbine- 
generators  operating  condensing.  The  prices  given  are  the 
averages  of  those  given  by  a  number  of  manufacturers  at  a  time 
when  the  cost  of  foundry  pig  iron  was  about  $20  per  ton.  It  is 
estimated  that  the  values  given  by  the  curves  will  be  changed 
roughly  about  2  per  cent,  for  a  variation  of  $i  in  the  price  of 
foundry  pig  iron. 

Unless  some  such  standard  of  values  is  given  such  results  can 
be  of  little  value  a  very  short  time  after  the  curves  are  prepared. 


292 


THE  STEAM  TURBINE 


Non-condensing  turbines  cost  about  5  per  cent,  less  than  con- 
densing machines.  Prices  of  25-cycle  and  6o-cycle  generators 
are  usually  about  the  same.  Prices  do  not  include  charges  for 
freight  and  erection,  which  in  the  eastern  and  middle  western 
states  are  about  $i  to  $1.50  per  kilowatt. 


50  100  150         200          250          300 

1,000       2,000       3,000      4,000      5,000       6,000 
Rated  Full  Load      Kw. 


FIG.  206.     Curves  of  the  Approximate  Price  of  Steam  Turbine-Generators  per 
Kilowatt  of  the  Rated  Full  Load  Output  of  the  Generator. 


Mr.  W.  C.  Gottshall,  who  has  very  carefully  investigated  power 
plant  economics,  has  collected  the  data  on  the  following  page, 
published  in  1903,  regarding  the  probable  maximum  and  mini- 
mum costs  per  rated  kilowatt  installed  of  a  power  plant  equip- 
ment of  about  10,000  kilowatts  capacity.* 

High-grade  power  stations  of  from  5000  to  10,000  kilowatts 
capacity  with  thoroughly  modern  equipments  cost  usually  from 
$100  to  $125  per  kilowatt.  In  a  few  very  large  stations  with 
high-grade  equipment  the  cost  has  been  about  $60  per  kilowatt 
installed;  but  for  stations  under  10,000  kilowatts'  capacity  the 
cost  is  rarely  below  $90  per  kilowatt. 

A  building  of  modern,  factory  type  of  construction  (one  story 
—  steel  and  glass)  costs  about  $i  per  square  foot  of  floor  space. 

*  Gottshall,  Street  Railway  Economics. 


STEAM  TURBINE  ECONOMICS 


293 


COST  OF  A  STEAM   POWER  HOUSE  AND  EQUIPMENT. 


. 

Costs  per  Rated  Kilowatt 
Installed. 

Maximum. 

Minimum. 

Boilers  and  settings  

$l  7  .  oo 
3.00 
4-50 
6.00 
1.50 

12.00 
2.00 
I.  00 
32.00 
21.00 
IO.OO 
4.00 

6.00 
3.00 

3-5° 
15.00 

2.OO 

$9.00 
2.50 
2.50 
2.OO 
I.  00 
4.00 
I  .00 
I.  00 

20.00 

18.00 

2.OO 
1.50 
3.00 
2.00 
•50 

8.00 

I.  00 

Stokers 

"Economizers 

Coal  conveyors  and  bunkers 

Ash  conveyor              

Piping  and  covering      

Feed-water  heater  

Feed  pumps  

Engines  or  turbines* 

Generators 

Condensers  including  pumpst 

Switchboard        ..... 

Power-house  cables  and  conduits        

Incidentals  (as  concrete  floor  and  traveling  crane)  .  .  . 
Foundations  for  machinervj  

Buildings 

Chimneys  and  flues 

Total  cost  including  10  per  cent,  for  engineer- 
ing supervision  and  contingencies  (nearly)  .  .  . 

$158.00    ' 

$87.00 

*  This  item  is  about  right  for  reciprocating  engines  and  turbines  in  1903,  when  these  data  were  pub- 
lished. Fig.  206  shows,  however,  that  the  minimum  cost  for  this  item  is  about  the  same  as  the  present 
cost  of  first-class  turbines  and  generators. 

t  The  cost  of  condensers  is  not  included  in  GottshalPs  data.  Prices  given  here  are  those  given 
by  J.  R.  Bibbins  (Report  American  Si.  Ry.  Assn.,  1004)  for  a  plant  operating  at  26  inches  vacuum. 
He  estimates  that  the  cost  of  a  plant  for  28  inches  vacuum  is  60  per  cent,  greater.  Bibbins'  values 
may  be  tabulated  as  follows: 

COST  OF  CONDENSING  PLANT  PER  RATED  KILOWATT. 


Inches  Vacuum. 


26 

28 

29 

Barometric  condenser  

$6  to  $7  .50 

$0-SO  tO  $12 

$12   tO  $15 

Surface   condenser   (including   centrifugal   lift 
pump,  air  cooler,  single-cylinder  dry  vacuum 
pump,  and  centrifugal  circulating  pump;  
Surface  condenser  (including  wet  vacuum  pump 
and  centrifugal  circulating  pump)  

$7.50  to  $10 

$12   tO  $l6 

$12  to  $16 

$15  tO  $20 

$15  to  $20 

Ejector  condenser  

$2           tO  $2.5O 

$3  to  $4 

$4  to  $5 

$  Engine  and  generator  foundations  cost  from  $1.00  to  $3.50  per  kilowatt  capacity.  Foundations 
for  turbine-generators  cost  as  a  rule  about  one-fourth  as  much,  usually  30  to  40  cents  per  kilowatt 
capacity  on  fairly  good  sub-soil. 


294  THE   STEAM   TURBINE 

Itemized  costs  of  the  85oo-kilowatt  power  plant  of  the  Fort 
Wayne  and  Wabash  Valley  Railway  Company  at  Fort  Wayne, 
Ind.,  are  given  by  Bibbins  as  follows: 

Substation  apparatus  and  buildings  are,  of  course,  not 
included.  Drawings  and  a  photograph  of  this  station  are  shown 
in  Figs.  199  and  200. 

The  double-deck  arrangement  and  the  installation  of  baro- 
metric condensers  designed  for  a  moderate  vacuum  make  the. 
first  cost  of  this  station  very  low. 

Dollars  per 
Kilowatt. 

Building:    Including  general  concrete  and  steel  work,  coal  bun- 
ker, smoke  flue,  condenser  pit,  coal  storage  pit,  etc IO-97 

Boiler  plant:    Including    boilers,    superheaters,    stokers,    piping, 

pumps,  heaters,  settings,  breechings,  and  tank 13 .92 

Generating  plant:    Including  turbines,  generators,  exciters,  cables, 

switchboards,  transformers,  and  ventilating  ducts 3°-55 

Condenser  plant:    Including    condensers,    pumps,    piping,    free 

exhausts,  water  tunnels,  and  intake  screen 3-9& 

Coal-handling  plant:  Including  gauntree  crane,  crusher  motors, 

and  track o .  94 

Erection,  superintendence,  and  engineering 5-94 

Total,  excluding  property  and  siding 66. 25 

The  costs  are  based  upon  the  following  assumptions: 

Per  Cent. 

(a)  Bond  interest  and  taxes 7 

(b)  Sinking  fund,  equivalent  to  6.43  per  cent,  depreciation 4.2 

(c)  Total  fixed  charges  on  capital  cost 11.2 

Depreciation  determined  by  summing  the  depreciation  on  the 
several  parts  of  the  plant  as  follows:  building,  3  per  cent.; 
boiler  plant  and  coal-handling  apparatus,  10  per  cent.;  condens- 
ing plant,  6  per  cent.;  generating  plant,  7.5  per  cent.;  general 
average,  6.43  per  cent. 

In  calculating  the  cost  of  an  electric  power  plant  it  is  necessary 
to  consider  the  probable  life  of  the  plant,  so  as  to  make  correct 
yearly  reductions  for  depreciation.  Often  this  is  more  or  less 
of  guess-work  on  the  part  of  the  constructor  or  owner,  and  for 


STEAM  TURBINE  ECONOMICS 


295 


this  reason  the  following  table  from  a  recent  issue  of  Zeitschrift 
des  Vereines  deutscher  Ingcnieure  is  interesting.  The  figures 
given  are  those  used  by  two  English  public  corporations,  two 
English  engineers,  and  a  series  of  figures  taken  from  German 
technical  publications.  Conditions  in  America  are  somewhat 
different  from  those  of  Europe  and  the  depreciation  is  usually 
somewhat  greater. 

ESTIMATED    YEARS    OF    LIFE. 


Authority. 

Local 
Gov't 
Board. 

L.  Canby 
Council. 

Robert 
Hammond. 

J.  F.  C. 
Snell. 

German 
Publi- 
cations. 

Buildings  

•7Q 

CQ 

00 

60 

66 

Boilers  

ir 

2O 

20 

20 

T  P 

Steam  engines  

ir—  2C 

2O 

2O—  2s 

2C. 

1J 
2O 

Steam  turbines  

2O 

Gas  engines  

I  7 

Water  turbines  

22 

Dynamos  

2O 

2O 

2S 

2  C. 

Storage  batteries  .  . 

c—  7 

2O 

I  C, 

*a 

IO 

Transformers  

I  c 

2O 

I  r 

20 

Switchboards  

1C. 

2O 

2O 

2O—  2? 

T  ? 

Electric  cables  (conductors) 

I2-IC, 

I2—5O 

10 

I  ^—60 

Electric  meters 

IO 

Arc  lights  

7—  IO 

IO 

*3 

1C. 

Comparisons  have  been  made  by  L.  G.  French  of  the  cost 
of  two  sizes  of  turbine-generators  with  corresponding  recipro- 
cating engine  costs.  He  states  that  the  cost  of  a  75o-kilowatt 
turbine-generator  with  a  surface  condenser  (operating  vacuum 
not  given)  and  including  foundations  and  installation  charges 
was  $37  per  kilowatt.  A  similar  reciprocating  engine  plant  cost 
$40  per  kilowatt.  A  1 5oo-kilowatt  turbine-generator  with  a  similar 
condenser  equipment  cost  $30.20  per  kilowatt,  including  founda- 
tions and  installation,  while  a  reciprocating  engine  equipped 
similarly  cost  $32.40  per  kilowatt. 

*  Orrok  of  the  N.  Y.  Edison  Co.  states  that  the  company  has  had  40  small 
steam  turbines  of  the  impulse  type  in  service  for  five  years  with  practically  no 
expense  for  repairs. 


296  THE  STEAM  TURBINE 

It  is  generally  believed  by  engineers  who  have  done  recent 
work  in  the  equipment  of  large  new  steam  power  stations  that  it 
is  not  very  probable  that  large  reciprocating  engines  will  ever  again 
be  installed  to  develop  power  for  electrical  distribution.  One 
reason  is  that  turbine-driven  alternators  are  particularly  adapt- 
able for  parallel  operation.  The  low  first  cost  and  operating 
expenses  of  turbine-generator  units  as  well  as  the  saving  in  the 
cost  of  foundations  and  floor  space  are  also  very  important 
considerations.  A  manufacturer  of  very  large  sizes  of  both 
steam  turbines  and  reciprocating  engines  has  stated  that  a  large 
power  station  if  equipped  with  reciprocating  engines  instead  of 
steam  turbines  would  cost  at  least  from  35  to  60  per  cent,  more 
than  a  turbine  station  of  the  same  capacity. 

The  following  interesting  tests  of  the  power  required  to  operate 
the  auxiliary  machinery  *  needed  for  a  Curtis  turbine  were 
reported  by  the  Turbine  Committee  of  the  National  Electric 
Light  Association  in  1905.  The  data  apply  to  the  auxiliaries  of 

*  "  The  quantity  of  circulating  water  required  for  high- vacuum  condensing 
plants  must  be  increased  from  the  old  standard  of  from  25  to  30  pounds  to  from  40 
to  60  pounds  of  water  per  pound  of  steam  condensed  for  moderate  temperatures, 
and  from  60  to  100  pounds  of  water  per  pound  of  steam  when  used  at  the  higher 
temperatures  common  to  cooling  tower  practice.  In  cases  of  excessive  head  or 
quantity  of  circulating  water  the  bulk  of  the  power  required  by  auxiliaries  is  due  to 
the  circulating  pump.  The  range  of  power  for  this  purpose  varies  so  widely  that 
the  older  method  of  assuming  a  given  type  of  plant  requiring  5,  10,  or  15  per  cent, 
of  the  total  steam  consumption  to  drive  auxiliaries  is  entirely  in  error  without  an 
accompanying  statement  denning  conditions  under  which  circulating  water  is 
pumped.  The  power  to  drive  the  air  pump  is  dependent  somewhat  upon  the 
vacuum,  but  particularly  upon  the  air  leakage  into  the  condensing  system.  It  was 
for  some  time  assumed  that  the  work  of  the  air  pump  corresponded  to  removing 
the  air  which  entered  the  boiler  in  solution  in  feed-water.  As  a  matter  of  fact, 
handling  the  air  in  solution  is  the  smallest  portion  of  work  done  by  an  air  pump,  the 
leakage  through  piping,  pipe  joints,  pores  of  castings,  stuffing-boxes,  etc.,  imposing 
the  greatest  duty,  the  total  quantity  of  air  to  be  handled  ranging  from  ten  or  fifteen 
to  thirty  or  forty  times  the  air  dissolved  in  ordinary  water.  The  actual  power  to 
drive  the  air  pump  should  in  good  practice  be  less  than  .00018  indicated  horsepower 
in  the  air  pump  cylinder  per  pound  of  exhaust  steam  per  hour.  As  the  amount  of 
power  necessary  to  drive  the  air  pump  is  a  comparatively  small  portion  of  the  total 
power  for  auxiliaries  a  slight  error  in  this  quantity  will  not  largely  affect  the  final 
result."  —  C.  C.  Moore,  Journal  of  Electricity,  Power,  and  Gas,  March,  1905. 


STEAM  TURBINE  ECONOMICS 


297 


one  of  the  5ooo-kilowatt  turbine-generators  of  the  Boston  Edison 
Company. 


Test   i. 

Test    2. 

Test  3. 

Kilowatts  on  turbine  

27I3- 
28.4 

3410. 
28.7 

4758. 
28.6 

Barometer 

20  .  ^3 

20.  CK 

20.06 

Boiler  feed  pump  

E 
13.0 

orsepower  Usec 
27.7 

1. 
27.4 

Circulating  pump 

60    I 

60.  I 

60.  1 

Dry  vacuum  pump  

24.  T> 

23.2 

23.8 

Step  bearing  pump                      .... 

6.4 

5.8 

C.6 

Wet  vacuum  pump              

8.6 

Q.  2 

9.8 

Totals    

122  .  3 

1^1  . 

135.7 

Per  cent,  power  of  auxiliaries  to  power 
of  turbine 

7    4 

2    Q 

2.  I 

Per  cent,  water  used  by  auxiliaries  to 
that  used  by  turbine    

8.4 

7-4 

c  .  7 

PRACTICAL   DESIGNING    OF   POWER   STATIONS. 

Special  Fields  for  Steam  Turbines  and  for  Reciprocating  Engines. 

Some  space  will  be  given  here  to  the  economic  considerations 
entering  into  the  design  of  a  modern  power  station  intended  for 
electric  distribution  of  power.  There  can  be  no  doubt  as  to  the 
status  of  the  steam  turbine  in  comparison  with  reciprocating 
engines  for  the  generation  of  electrical  energy.  Practically  all 
the  recently  designed  power  stations  for  electric  services  are 
equipped  with  steam  turbine-generators,  and  in  some  of  the 
older  stations  built  originally  for  engine-driven  units  it  is  not 
unusual  to  see  turbines  installed  to  increase  the  capacity.* 
There  is  a  marked  contrast  between  a  power  station  equipped 
with  reciprocating  engines  and  one  that  is  turbine-driven.  The 
large  and  heavy  frames,  ponderous  moving  parts,  and  the 
large  generators  of  reciprocating  engine  plants  cannot  be  made 
to  compete  successfully  with  the  smaller,  more  compact,  and 
cheaper  turbine  units.  But,  on  the  other  hand,  reciprocating 

*  W.  C.  L.  Eglin,  Report  National  Electric  Light  Association,  1906. 


298 


THE  STEAM  TURBINE 


engines  similar  to  the  Corliss  type  have  a  field  which  for  a  num- 
ber of  years  probably,  the  steam  turbine  cannot  enter  success- 
fully. For  irregular  loads,  suddenly  applied,  like  those  of  rolling 


FIG.   195.     Comparative  Floor  Space  Required  for  Curtis  Turbines. 

mills  and  mine  hoists  the  reciprocating  engine  has  advantages 
over  the  steam  turbine,  except  perhaps  when  the  turbine  is  used 
in  connection  with  an  electric  drive.  Because  reciprocating 


STEAM  TURBINE   ECONOMICS  299 

water  pumps  and  air  compressors  are  more  efficient,  at  least  up 
to  the  present  time,  than  centrifugal  pumps  and  compressors, 
reciprocating  engines  are  invariably  installed  in  waterworks  and 
compressing  plants.  In  cotton  and  woolen  mills  and  shops, 
where  the  power  is  transmitted  by  belts,  shafting,  and  ropes 
instead  of  by  electrical  methods,  the  reciprocating  engine  because 
of  its  slower  speed  is  generally  preferable. 

Stated  in  a  few  words,  the  steam  turbine  is  unrivaled  by  steam 
reciprocating  engines  for  driving  apparatus  which  can  be  operated 
efficiently  at  a  high  speed  so  that  a  direct-connected  unit  can  be 
made. 

Parallel  operation  of  alternators  is  greatly  facilitated  when  they 
are  driven  by  turbines  rather  than  by  reciprocating  engines.  There 
are  always  difficulties  when  reciprocating  motion  is  to  be  con- 
verted into  synchronous  motion.  Besides  the  advantages  of  a 
uniform  turning  moment  which  makes  possible  such  close  speed 
regulation  that  it  is  possible  to  operate  railway,  power,  and 
lighting  circuits  from  one  turbine,  because  of  its  high  speed  it 
produces  a  more  powerful  regulating  force  without  the  use  of  a 
fly-wheel  than  that  of  any  engine-driven  units  of  the  same  capacity. 
Where  a  steam  turbine  is  installed  in  a  plant  with  piston  engines 
or  water-wheels  its  inertia  or  fly-wheel  effect  has  a  steadying 
effect  on  the  whole  system.  As  an  example  of  this  inertia  effect 
it  is  stated  *  that  a  35oo-kilowatt  Curtis  turbine  and  generator 
has  at  the  rated  speed  (750  r.p.m.)  a  "  storage  energy  "  of 
30,600,000  foot-pounds  which  is  sufficient  to  enable  the  machine 
to  carry  at  any  load  an  additional  load  equal  to  the  full  rating 
for  about  .75  of  a  second  with  a  drop  in  speed  of  only  3 
per  cent,  and  without  additional  steam.  This  machine  could 
carry  a  momentary  increase  of  load  of  half  the  rating  for  1.5 
seconds. 

Floor  Space  for  Power  Plants.  "  Compactness  "  expresses  well 
the  primary  requisite  for  the  economical  design  of  modern  power 
stations.  The  small  space  occupied  by  a  Curtis  steam  turbine 
compared  with  that  required  for  a  reciprocating  (Corliss)  engine 

*  A.  H.  Kruesi,  Proc.  American  Street  and  Inlerurban  Railway  Association,  1907. 


300 


THE  STEAM  TURBINE 


of  the  same  capacity  is  well  shown  by  Fig.  195.  Floor  space 
occupied  by  Westinghouse  turbine-generators  is  given  by  the 
curves  in  Fig.  196,  showing  the  number  of  square  feet  occupied 
per  kilowatt  or  per  brake  horsepower.  Comparisons  of  the 
space  required  for  power  units  are  of  little  value,  however,  unless 
the  space  for  the  condensing  apparatus  and  auxiliary  machinery 
is  also  considered.  It  is  probably  fair  to  assume  that  for  the 


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FIG.  196.    Floor  Space  Required  for  Westinghouse  Turbines. 


conditions  where  a  reciprocating  engine  would  be  operated  at  26 
inches  vacuum  condensers  for  a  turbine  plant  would  be  designed 
for  28  inches  vacuum.  Now  the  volume  of  steam  at  26  inches 
vacuum  is  very  nearly  half  that  at  28  inches.  When  surface 
condensers  are  used,  therefore,  the  very  great  increase  in  the  size 
of  the  condenser  equipment  for  turbine  plants  is  very  obvious. 
For  this  reason  there  has  been  a  tendency  in  recent  years  to 
install  barometric  or  the  open  type  of  jet  condenser  for  steam 
turbines. 

A  very  recent  installation  of  Westinghouse-Parsons  turbines, 
barometric   condensers,    and    Stirling   boilers   is   illustrated   in 


STEAM  TURBINE  ECONOMICS 


301 


-- 

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iL.- 

y  £__._ =-!^ss^=n^=>.-=s.?s=sa=irt^£llJJ.J 


FIG.  197.     "Double-Deck"  Design  of  Power  House  Equipped  with  Horizontal 

Turbines. 

Fig.  197.*    The  important  features  of  this  design  are  the  placing 
of  the  turbines  above  the  boilers  and  condensers,  the  use  of 

*  J.  R.  Bibbins,  Ttans.  Am.  Inst.  Elect.  Eng.,  1908. 


302  THE  STEAM  TURBINE 

barometric  condensers,  and  the  low  total  cost  of  power  house  and 
equipment.  It  is  probably  one  of  the  most  compact  arrange- 
ments possible  in  a  steam  turbine  plant  consistent  with  high 
economy  in  operation. 

In  this  "double-deck"  design  the  horizontal  turbine  and 
barometric  condensers  are  at  their  best  advantage  as  regards 
compactness  and  efficiency.*  The  connections  between  the 
turbines  and  the  condensers  are  short  and  direct,  which  obviates 
the  losses  occurring  where  there  are  bends  in  these  connections, 
and  the  cost  of  large  exhaust  piping  is  saved.  The  atmospheric 
relief  valve  of  the  turbine  is  placed  between  the  floor  girders,  so 
that  it  was  possible  to  make  the  distance  between  the  floor  level 
and  the  condenser  head  only  2.5  feet.  For  stations  not  at  tide- 
water, turbine  plants  are  usually  operated  at  a  moderate  vacuum 
of  between  27  and  28  inches.  Barometric  condensers  are  now 
being  made  to  maintain  this  vacuum  without  the  use  of  auxiliary 
dry-air  pumps. 

With  surface  condensers  by  far  the  most  compact  arrangement 
is  obtained  by  installing  Curtis  vertical  turbines  with  a  condenser 
base.  By  this  arrangement  a  very  direct  connection  between 
the  turbine  and  the  condenser  is  secured;  but  in  places  where 
there  is  likely  to  be  trouble  with  leaky  tubes  most  engineers  will 
prefer  a  condenser  separate  from  the  turbine. 

Drawings  showing  the  cross-section  and  plan  of  a  design  for 
horizontal  turbines,  Babcock  &  Wilcox  boilers,  and  barometric 
condensers  are  shown  in  Fig.  199.  An  exterior  view  of  the 
same  station  showing  the  coal-handling  equipment  is  illustrated 
by  Fig.  200. 

Plan  and  elevation  of  a  power  station  with  the  turbines  and 
boilers  on  approximately  the  same  floor  level  are  represented  by 
the  drawings  in  Figs.  201  and  202.  The  first  of  these  figures  is 
particularly  interesting  because  it  shows  very  clearly  the  arrange- 

*  A  similar  "double-deck"  arrangement  has  been  proposed  for  power  plants 
operated  by  horizontal  gas  engines  and  producers.  In  such  a  design,  where  pro- 
ducers, scrubbers,  and  all  auxiliaries  are  placed  on  the  second  floor,  it  has  been 
shown  that  the  ground-floor  area  was  only  2.25  square  feet  per  kilowatt. 


STEAM  TURBINE  ECONOMICS 


303 


TYPICAL     FLOOR      STRUCTUSC      FT.  tV/CTNC     STATION 


LONGITUDINAL   FLOOR   ELEVATION 


TRANSVERSE   FLOOR   FLEVATION 

FIG.  199.    Longitudinal  and  Transverse  Sections  of  Power  Station. 


304 


THE  STEAM  TURBINE 


ment  of  the  auxiliaries  in  stations  equipped  with  surface  con- 
densers. Fig.  203  is  intended  to  show  particularly  the  piping 
arrangements  for  a  typical  power  station  having  the  turbines 
and  auxiliary  equipment  in  a  room  adjoining  the  boiler  room. 


FIG.   200.     View  of  the  Power  House  Shown  in  Fig.  199. 


OILING   SYSTEMS   FOR   STEAM   TURBINES. 

A  perfect  oiling  system  is  obviously  a  necessity  for  any 
machinery  operating  at  a  high  speed.  The  efficiency  of  turbines 
of  the  Parsons  type  depends  largely  on  the  smallness  of  the  radial 
clearances  between  the  rotor  and  the  casing.  Now  if  there  is 
any  displacement  of  the  rotor  with  respect  to  the  casing,  caused, 
for  example,  by  the  melting  of  the  white  metal  in  one  of  the  main 
bearings,  the  blading  might  be  entirely  torn  or  "stripped,"  and 
the  turbine  would  probably  be  out  of  service  for  several  weeks. 
In  Curtis  turbines  with  vertical  shafts,  on  the  other  hand,  very 
serious  results  might  occur  if  the  flow  of  oil  to  the  step  bearing 
should  be  interrupted. 


STEAM  TURBINE  ECONOMICS 


305 


FIG.  201.     Plan  and  Elevation  of  Turbine  and  Condensing  Plant. 


306 


THE  STEAM  TURBINE 


STEAM  TURBINE  ECONOMICS 


307 


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308  THE  STEAM  TURBINE 

There  are  two  usual  methods  of  lubrication  for  steam  turbines: 
(i)  the  central  system  and  (2)  the  single  unit  system. 

In  the  case  of  the  central  system  an  oil  tank  is  placed  at  a  high 
point  in  the  building  and  the  oil  flows  through  pipes  by  gravity 
to  the  bearings  of  the  turbine.  By  means  of  a  "parallel"  system 
of  piping  any  number  of  turbines  can  be  supplied  from  one  oil 
tank.  The  oil  leaving  the  bearings  flows  into  a  suitable  filtering 
apparatus  provided  with  cooling  coils  from  which  it  is  pumped 
back  to  the  main  supply  tank.  The  chief  objection  to  this  system 
Is  the  danger  of  a  total  shut-down  of  the  oiling  system  caused  by 
a  poor  joint  or  a  broken  pipe  between  the  supply  tank  and  the 
turbine  bearings. 

The  alternate  system,  in  which  each  turbine  has  its  own  oil 
supply  and  pump,  has  the  advantage  in  that  it  assists  in  reducing 
the  risk  of  a  total  shut-down  of  the  plant  to  a  minimum,  and  if 
the  oil  is  spoiled  in  one  turbine,  due  to  being  mixed  with  water 
or  being  overheated,  the  entire  supply  of  the  station  is  not 
ruined. 

Until  recently  nearly  all  manufacturers  of  Parsons  turbines 
supplied  their  machines  with  plunger  reciprocating  oil  pumps. 
In  this  respect  an  innovation  has  been  introduced  in  Westing- 
house  turbines  by  the  use  of  a  rotary  oil  pump  shown  in  Fig.  204. 
In  the  drawings  shown  here  there  are  two  sectional  views  of  the 
pump.  A  worm  gear  on  the  turbine  shaft  transmits  power  to 
the  pump  by  means  of  the  gear  wheel  10.  The  direction  of 
rotation  of  the  shaft  and  of  the  flow  of  oil  is  shown  by  arrows  in 
the  sections.  The  pump  cylinder  and  its  rotor  are  not  con- 
centric, and  metal  strips,  backed  by  springs,  are  inserted  into 
slots  in  the  rotor.  These  strips  are  forced  out  by  the  springs  to 
touch  the  inside  of  the  pump  cylinder  in  every  position,  so  as  to 
form  pockets  into  which  the  oil  enters  on  one  side  and  is  dis- 
charged from  the  other  side.  Similar  rotary  pumps  are  very 
generally  used  for  all  kinds  of  engineering  services. 

A  suitable  oiling  system  for  a  Curtis  turbine  (including  the 
step  bearing)  is  well  illustrated  diagrammatically  by  Fig.  205. 
A  large  storage  tank,  shown  at  the  right-hand  side  of  the  figure, 


STEAM  TURBINE  ECONOMICS 


309 


is  fitted  with  suitable  straining  devices  and  a  cooling  coil.  It  is 
usually  located  low  enough  to  receive  oil  by  gravity  from  all 
parts  requiring  lubrication.  Oil  from  this  tank  flows  to  a  pump 
from  which  it  is  discharged  at  a  pressure  about  25  per  cent, 
greater  than  that  required  to  sustain  the  weight  of  the  shaft  and 
wheels  on  the  step  bearing.  A  baffler  in  the  form  of  an  adjustable 


12 


FIG.  204.     Westinghouse  Oil  Pump. 

spiral  inserted  in  the  pipe  leading  to  the  step  bearing  serves  to 
regulate  the  oil  supply.  Another  line  of  piping  is  provided  for 
oiling  the  upper  parts  of  the  turbine.*  This  line  of  piping  is 
provided  with  a  reducing  valve  and  an  air  chamber  partly  filled 
with  compressed  air  to  maintain  a  constant  pressure  necessary 
for  the  hydraulic  motor  operating  the  valve  mechanism.  Drain 

*  Oil  pressure  on  the  upper  bearings  is  about  60  pounds  per  square  inch. 


THE  STEAM  TURBINE 


TO  5PRLMG   £QUALIZ£P 
OR  ACCUMULATOR 


FIG.  205.     Oiling  System  for  a  Curtis  Turbine. 


STEAM  TURBINE  ECONOMICS  311 

pipes  from  the  upper  bearings  and  from  the  hydraulic  motor  dis- 
charge into  a  common  receiver  in  which  the  streams  are  visible, 
so  that  the  oil  distribution  can  be  always  observed. 

At  some  point  in  the  high-pressure  system  adjacent  to  the  pump 
a  device  is  usually  installed  to  equalize  the  discharge  of  oil  from 
the  pump.  Ordinarily  Curtis  turbines  are  provided  with  a  small 
spring  accumulator  for  this  purpose,  except  for  cases  where 
weighted  storage  accumulators  are  to  be  installed.  A  storage 
accumulator  is  usually  recommended  for  large  power  stations. 
It  can  be  arranged  so  that  it  will  normally  remain  full,  but  will 
discharge  if  the  pressure  fails,  and  start  automatically  auxiliary 
pumps. 

Piping  for  Superheated  Steam.  Much  of  the  trouble  resulting 
from  the  use  of  superheated  steam  is  due  not  so  much  to  want  of 
strength  as  to  the  want  of  elasticity  in  the  parts  affected.  These 
troubles  are  due  particularly  to  the  unceasing  variations  in 
temperature  resulting  from  fluctuating  loads  rather  than  from 
high  temperatures.  As  it  is  possible  for  water  to  exist  in  the 
liquid  state  in  superheated  steam,  the  variations  in  temperature 
may  produce  a  spraying  of  highly  heated  surfaces,  which  greatly 
increases  these  difficulties.  Changes  in  the  design  of  pipe 
fittings,  valves,  boilers,  and  superheaters  should  be  made  to  allow 
for  this  abnormal  condition.  It  is  desirable  to  use  annealed  steel 
castings  in  place  of  cast-iron  for  fittings  and  valve  casings,  and 
the  use  of  copper  for  internal  parts  of  valves  and  gaskets  should 
be  avoided.  Low  velocities  in  steam  piping,  which  have  become 
customary  on  account  of  the  pulsating  flow  of  reciprocating 
steam  engines,  are  not  suitable  for  superheated  steam.  Since 
flexibility  is  so  important  a  consideration  in  piping  for  super- 
heated steam,  it  is  necessary  to  use  comparatively  small  sizes  of 
pipes  and  fittings. 

In  Curtis  turbines,  Kruesi  states,  a  velocity  of  at  least  140  feet 
per  second  (about  8500  feet  per  minute)  is  desirable  for  dry  sat- 
urated steam.  Now  if  the  steam  is  superheated  100  degrees  F. 
the  volume  is  increased  15  per  cent.,  but  "the  velocity  in  the 
pipes  will  be  substantially  the  same  on  account  of  the  reduction 


312  THE  STEAM  TURBINE 

in  the  steam  consumption  of  the  turbine.57  Although  this 
statement  is  not  quite  accurate  because  the  steam  consumption  of 
Curtis  turbines  is  usually  reduced  only  8  to  10  per  cent,  per  100 
degrees  F.  superheat,  it  is  an  important  observation  that  the  size 
of  piping  should  not  be  increased  in  proportion  to  the  increase  in 
volume  of  the  steam  due  to  superheating. 


CHAPTER    XV. 
STRESSES   IN  RINGS,   DRUMS,  AND  DISKS. 

Design  of  a  Bucket  Band  or  Ring.  A  ring  or  band  is  one  of  the 
simplest  means  of  fastening  together  a  number  of  separate  pieces 
attached  like  the  blades  of  a  turbine  wheel  to  the  circumference 
of  a  cylindrical  surface.  Such  bands  are  always  made  a  little 
wider  than  the  blades,  especially  at  the  side  where  the  steam 
enters,  so  that  the  edges  of  the  blades  may  not  be  easily  damaged 
in  transportation  and  from  insufficient  axial  clearances  when  the 
turbine  is  operated. 

These  bands  are  very  serviceable  in  taking  care  of  loose  buckets 
which  otherwise  would  be  troublesome.  The  band  serves  to 
bind  the  blades  together  as  a  whole,  making  the  blades  with  weak 
attachments  to  the  wheel  as  good  as  the  strongest.  The  band 
assists  in  making  a  row  of  blades  of  uniform  strength.* 

The  design  of  such  a  ring  revolving  at  high  speeds  should  be 
determined  by  careful  calculations;  but  the  theory  underlying 
the  design  of  such  a  ring  serves  also  for  the  design  of  turbine 
drums  and  disks. 

Centrifugal  forces  more  than  any  other  considerations  deter- 
mine the  design  of  a  blade  ring  or  band  for  strength.  These 
forces  produce,  of  course,  tension  and  a  resulting  expansion  of  the 
ring  —  both  of  significant  importance. 

The  centrifugal  force   (CF)  in  any  sector  (W  pounds)  of  a 

*  In  a  Parsons  type  it  cannot  be  assumed,  however,  that  because  the  blades  can 
be  made  stiffer  by  the  use  of  a  band  or  shroud  ring  it  is  possible  to  reduce  radial 
clearances  below  the  normal  amount  and  at  the  same  time  reduce  leakage  around 
the  blades.  There  is  reason  for  believing  that  radial  clearances  should  be  increased 
for  satisfactory  operation  when  the  "band  "  construction  is  used  unless  the  relative 
expansion  of  the  metals  in  the  ring,  blades,  drum,  and  casing  is  very  carefully 
adjusted. 

3*3 


THE  STEAM  TURBINE 


freely  rotating  ring  of  radius  r  inches,  velocity  V  feet  per  second, 
with  an  angle  6  subtended  by  the  sector,  is 


CF  = 


where  g  is  the  acceleration  due  to  gravity. 

This  centrifugal  force  tending  to  expand  the  ring  by  increasing 
its  circumferential  dimensions 
sets  up  stresses  which,  for 
the  purposes  of  calculation, 
may  be  represented  by  tan- 
gential forces  at  the  ends  of 
the  sector.  These  forces  are 
necessarily  equal  for  equilib- 
rium and  are  shown  as  T  and 
T  in  Fig.  208.  If  the  breadth 
of  the  sector  is  represented 
by  m  inches  and  the  radial 
thickness  by  n  inches,  then 
the  area  of  the  section  over 
which  this  stress  is  distributed  Fl0'  2°8'  Fo'ces '°  *?  Blade  Band 

or  Shroud  Ring. 

is  mn  square  inches;   and  if 

S  is  the  unit  tensile  stress  in  pounds  per  square  inch,  each  tan- 
gential force  is  expressed  by 

T  =  mnS. 

This  force  T  on  the  section  is  tangential,  and  since  the  radial 
centrifugal  force  (CF)  must  be  equilibrated  by  an  equivalent 
radial  force  TO  *  or  for  equilibrium 

CF  =  T0  f, 

WV2 

—  =mnS0- 


*  This  relation  is  obvious  from  the  geometry  of  the  figure.  It  is,  of  course, 
not  quite  accurate,  but  very  nearly  correct  for  small  values  of  6. 

f  It  cannot  be  assumed  that  at  the  moment  of  rupture  the  stress  will  be  dis- 
tributed between  the  two  sections.  The  assumption  made  in  the  equations  is, 
however,  very  much  on  the  safe  side. 


STRESSES  IN  RINGS,  DRUMS,  AND   DISKS  3 1 5 

Now  if  z  is  the  weight  in  pounds  of  a  cubic  inch  of  the  material 
of  the  ring,  the  length  of  the  sector  (Fig.  208)  is  rO  inches;  then 


(29) 


This  equation  shows  that  the  unit  stress  in  a  blade  ring  or  band 
depends  only  on  the  weight  of  the  material  and  on  the  peripheral 
velocity.  The  last  equation  can  also  be  expressed  in  another 
form,  remembering  that 

y_  3.1416  dN 

60  X    12   ' 

where  d  is  the  diameter  in  inches  to  the  central  line  of  the  ring 
and  N  is  the  number  of  revolutions  per  minute.  Then  if  we 
make  the  approximation  of  TT*  =  10,  we  have 

S  = 

g  x  4320 

^2Tvr2 

(30) 

Equations  (29)  and  (30)  are  generally  used  for  the  design  of 
shroudings  and  overhanging  rims.  When  such  rings  are  per- 
forated with  small  holes  for  the  riveting  of  blades,  bending  and 
shear  are  produced.  The  stresses  due  to  this  bending  and 
shear  are,  however,  small  and  do  not  in  practical  cases  often 
exceed  400  pounds  per  square  inch. 

Sometimes  rings  called  "segments"  (Fig.  115)  are  put  on  the 
edge  of  wheel  disks  and  the  blades  are  attached  to  them.  In  a 
construction  of  this  kind  the  ring  must  not  only  restrain  the  cen- 
trifugal force  due  to  its  own  weight,  but  also  part  of  that  from 
the  weight  of  the  blades  if  they  are  not  tightly  fitted. 


3l6  THE  STEAM  TURBINE 

If  the  following  symbols  are  assumed: 

r0  =  radius  to  center  line  of  blades,  in  inches, 
d0  =  diameter  to  center  line  of  blades,  in  inches, 
w    =  weight  of  blades  in  pounds  per  foot  of  length  of  the 
circumference    measured    to    the    center   line    of   the 
blade  ring, 

rO  =  length  of  a  short  segment  of  the  blade  ring  to  be  calcu- 
lated, in  inches, 

V0  =  peripheral  velocity  of  blades,  in  feet  per  second, 
W0  =  weight  of  the  blades,  in  pounds,  of  a  segment  r0#  inches 

wrftfl 
long  or  — "-  pounds, 

then  the  centrifugal  force  at  the  blade  ring  due  to  the  weight  of 
the  blades  alone   is 

2       wr00V02       w0V2 


Then  if  T0  is  the  tangential  force  in  the  blade  ring  due  to  the 
weight  of  the  blades,  and  S0  is  the  corresponding  unit  tensile 
stress  in  pounds  per  square  inch, 

C0  =  T0#  =  mnS00  and 

=  mnS00, 
g 

S0  =  ^, 
gmn 

or 

g  W7r2d02N2  wd0W  wd0W  ,, 

518,400  gmn      51,840  gmn        1,669,200  mn  * 

then  the  total  stress  St  due  to  the  weight  of  the  ring  and  of  the 
blades  is 

zd2N2  wd02N* 

139,100       1,669,200  mn 


STRESSES  IN  RINGS,  DRUMS,  AND  DISKS  317 

If  a  blade  band  or  shroud  is  made  in  a  solid  ring  and  is  shrunk 
on  the  outside  of  the  blades,  as  is  sometimes  the  case,  then  the 
elongation  of  the  ring  due  to  the  centrifugal  stresses  must  be 
allowed  for.  In  other  words,  the  ring  must  be  made  small 
enough  so  that  there  will  be  a  tight  fit  at  the  highest  speed  that 
will  ever  be  attained.* 

Design  of  Drums  for  the  Rotors  of  Reaction  Turbines.  The 
blades  of  steam  turbines  are,  as  a  rule,  fastened  to  a  cylindrical 
drum  or  to  one  or  more  disks.  The  drum  construction  is  used 
where  there  is  a  large  number  of  stages  with  a  small  drop  of 
pressure  between  the  successive  stages  and  usually  comparatively 
low  peripheral  speeds.  Thus  the  rotor  of  a  Parsons  type  is 
made  up  of  a  number  of  drums  of  different  diameters,  increasing 
in  size  toward  the  low-pressure  end.  The  drum  diameters  are 
determined  by  the  blade  speed  which  is  selected  by  the  designer 
to  give  approximately  the  best  efficiency  for  the  velocity  of  the 
steam  in  the  stages  of  each  section  of  the  rotor. 

Calculations  to  determine  the  thickness  of  a  section  of  the  drum 
are  the  same  in  principle  as  for  a  blade  ring  as  explained  in  the 
preceding  paragraphs. 

The  thickness  of  the  drum  shell  is  most  simply  determined  by 
making  calculations  in  the  following  order: 

(i)  Calculate  the  stress  in  the  cylindrical  shell  of  the  drum 
due  to  its  own  weight  by  equation  (30).  This  stress  can  be 
determined  immediately  because  it  is  independent  of  all  dimen- 
sions of  the  drum  except  the  diameter  of  the  shell  at  its  center 
line.  It  is  assumed,  of  course,  that  before  the  thickness  of  the 

*  If  5   =  elongation  per  inch  of  length, 

S  =  the  unit  stress  Ibs.  per  sq.  inch  in  the  ring  at  the  maximum  speed 

attained, 
E  =  modulus  of  elasticity  in  Ibs.  per  sq.  inch, 

then        s  =  — ,  and  the   total  elongation  of    circumference  is    •          inches. 
This  means  then  that  the  circumference  of  the  ring  must  be  made  — - —   inches 

tL, 

smaller  than  if  not  subjected  to  centrifugal  stress.  A  very  common  construction  is, 
however,  that  of  making  the  ring  in  segments  of  about  2  feet  in  length  and  riveting 
the  blades  to  these  segments. 


3l8  THE  STEAM  TURBINE 

metal  for  the  drum  is  to  be  determined,  the  blades  have  been 
designed  so  that  their  weight  can  be  calculated. 

(2)  Allowable  unit  tensile  stress  must  be  determined.     In  this 
connection  the  factors  to  be  considered  are  the  qualities  of  the 
material  to  be  used  (see  pages  337  and  338)  and  the  grade  of  work- 
manship that  is  available.     In  some  shops  in  Germany  where 
very  expert  workmen  can  be  secured  and  the  material  is  carefully 
selected  and  unusually  good,  a  factor  of  safety  as  low  as  three  is 
sometimes   used.     Manufacturers   of   De  Laval   turbine   wheels 
make  the   limiting  factor  from   four  to   five;   but   for  average 
American  practice  a  factor  of  safety  of  less  than  five  should  not 
be  considered.     If  nickel  steel  is  to  be  used  of  which  the  ultimate 
strength  is  say  120,000  pounds  per  square  inch,  with  a  factor  of 
safety  of  five,  the  allowable  total  stress  in  the  drum  shell  would  be 
24,000  pounds  per  square  inch.     Now  if  the  stress  due  to  its 
own  weight,  of  which  the  calculation  has  already  been  indicated, 
is  still  represented  by  the  symbol  S,  and  the  total  stress  allowable 
by  Sty  then  the  permissible  stress  resulting  from  the  weight  of 
the  blades  S0  is 

S0  =  St  -  S  =  24,000  -  S. 

(3)  The  thickness  of  the  drum  shell  can  now  be  calculated  by 

equation  (31).     Since  S0  is  now  determined  and  d0,  N,  andw*  are 
given  by  the  dimensions  required  for  the  design  of  the  blades, 
the  thickness  n  can  be  easily  calculated. 
Equation  (31)  can  be  written  in  the  form 


(33) 


Since  the  weight  of  the  blades  has  been  calculated  for  only  one 
row,  the  dimension  m  is  the  distance  between  the  center  lines  of 
successive  blade  rows  on  the  drum. 

*  Blades  made  of  bronze,  zinc,  copper,  or  similar  alloys  weigh  about  .30  pound 
per  cubic  inch,  and  steel  weighs  .28  pound  per  cubic  inch. 


STRESSES  IX  RINGS,  DRUMS,  AND  DISKS  319 

Example.  The  following  data  regarding  the  shell  of  a  section 
of  a  turbine  rotor  are  given  by  the  drawings  accompanying  the 
blade  design. 

Diameter  at  root  of  blades  (approximately  =  d) 25  inches 

Diameter  at  center  line  of  blades  (d0) 30  inches 

Revolutions  per  minute 2000 

Weight  of  blades  in  one  row,  per  foot  (w) 5  pounds 

Weight  of  a  cubic  inch  of  material  of  shell 28  pound 

Distance  between  center  lines  of  successive  rows  of  blades  of 

the  drum 3  inches 

The  stress  in  the  shell  (S)  due  to  its  own  weight,  by  equation 
(30),  is 

.28(25)2   (2000 )2  .       , 

S    =  -  -*-  =  ^030  pounds  per  square  inch. 

139,100 

S0  =  24,000  —  5030  =  18,970  pounds  per  square  inch. 

5  (so)2  (2000 )2  .     , 

mn  =  — ^Vs7   '    v —  *-f =  .57  square  inch. 

1,669,200  X  18,970 

But    m  =  3  inches;  then 

n  =  -57  -*-  3  =  ^9  inch- 

The  sections  of  the  rotor  are  usually  supported  on  disks  attached 
to  the  shaft.  In  another  paragraph  relating  to  the  design  of 
disks,  the  strength  of  such  forms  will  be  discussed.  It  should  be 
remembered  that,  compared  with  impulse  turbines,  the  peripheral 
speed  is  always  kept  low.*  Drums  are  almost  always  used  for 
reaction  turbines,  and  separate  disks  or  wheels  for  impulse 
turbines. 

Fig.  209  is  an  exact  copy  of  the  shop  drawing  of  the  rotor  of  an 
Allis-Chalmers  (Parsons  type)  turbine.  It  consists  of  a  central 
cylinder  upon  which  rings  are  fitted  as  shown.  These  rings  are 
made  of  steel  and  are  forged  as  a  solid  ring.  The  webs  -are 
formed  by  cutting  away  the  superfluous  material  in  the  sides 
with  a  lathe.  In  this  type  of  rotor  the  central  cylinder  must  be 
made  of  sufficient  strength  to  resist  the  usual  torsional  stresses 

*  The  peripheral  velocity  of  drum  types  should  not  exceed  400  feet  per  second. 
Impulse  wheels,  however,  are  sometimes  designed  to  operate  at  1200  feet  per 
second. 


320 


THE  STEAM  TURBINE 


in  a  "  hollow  "  shaft.     The  construction  of  the  drums  of  typical 
Westinghouse  turbines  is  shown  in  Figs.  107,  109  and  184. 

In  impulse  turbines  where  all  the  expansion  of  the  steam  takes 
place  in  nozzles  placed  in  diaphragms,  or  partitions  between  the 
stages,  there  is  a  large  drop  in  pressure  between  any  two  stages, 
and  therefore  leakage  of  steam  between  the  stages  will  be  much 
greater  than  with  the  small  pressure  drop  in  the  reaction  type. 
The  fewer  number  of  stages  in  the  impulse  turbine  necessarily 


FIG.  209.     Section  of  the  Rotor  of  a  Parsons  Type  of  Reaction  Turbine. 

increases  the  velocity  of  the  steam  passing  through  the  blades 
and  at  the  same  time  the  most  economical  wheel  speed.  Within 
practical  limits,  wheel  speed  should  always  be  increased  with 
steam  velocity  in  good  designing. 

Stresses  at  Right  Angles  to  Each  Other.  To  determine  the 
stress  in  flat  disks  a  refinement  in  the  calculations  is  sometimes 
necessary  in  order  to  obtain  more  accurate  values  than  those 
secured  in  the  preceding  calculations  for  the  stresses  in  rings  and 
drums.  If,  for  example,  two  forces  R  and  T  act  at  right  angles 
to  each  other,  theoretical  conditions  of  elasticity  show  that  the 
maximum  stress  or  elongation  is  never  quite  equal  to  that  due 
to  either  of  the  two  forces  if  acting  alone.  In  other  words,  an 
elongation  in  the  direction  of  the  line  of  action  of  the  force  R 
produces  a  contraction  in  the  direction  of  the  force  T.*  Thus 

*  This  phenomenon  is  easily  observed  in  a  piece  of  india-rubber.  A  force  in 
one  direction  producing  an  elongation  will  produce  also  a  contraction  in  the 
direction  at  right  angles  to  the  greatest  elongation. 


STRESSES  IN  £INGS,  DRUMS,  AND  DISKS  321 

if  the  elongation  due  to  the  force  R  is  sr  per  unit  of  length  we  have 
the  relation 

s'* 

c      —  —  JK 

Sr  ~E 

where  Sr  is  the  stress  in  pounds  per  square  inch  and  E  is  the 
modulus  of  elasticity  of  the  material.  The  reduction  (snr)  of  the 
dimension  at  right  angles  or  normal  to  the  direction  of  the  force 
producing  the  elongation  is  proportional  to  the  force  itself  and 
also,  of  course,  to  the  stress.  Then 

kSr 

Snr  =  kSr  =  —  , 

hi 

where  k  is  a  constant  and  has  the  value  of  .3  for  metals  of  a 
homogeneous  structure,   such  as  are  usually  required  for  the 
manufacture  of  machines. 
The  force  T  in  the  same  way  produces  elongation 

-I  '••;- 

and  a  reduction  at  right  angles  (snt)  (in  direction  opposite  to  the 

elongation  due  to  Sr), 

_kS, 
'  E 

The  net  elongation  in  the  direction  of  the  force  R  is 


Also  the  net  elongation  in  the  direction  of  the  force  T  is 

St      kSt      1 

"S"r=E"¥=E(S'~'3Sr)- 

When  the  two  stresses  at  right  angles  are  nearly  equal,  as  in 
the  case  of  the  disk  now  under  consideration,  the  elongation  is, 
from  the  results  above,  only  .7  of  that  resulting  from  either  force 

*  See  Greene's  Structural  Mechanics,  pages  7  and  184;  Church's  Mechanics  of 
Engineering,  p.  203. 


322 


THE  STEAM  TURBINE 


acting  alone.  It  follows  also  that  when  the  stresses  are  nearly 
equal  the  stresses  which  are,  of  course,  proportional  to  defor- 
mations are  also  only  .7  of  that  calculated  from  only  one  of  the 
forces.  This  effect  of  forces  at  right  angles  to  each  other  will 
be  applied  in  the  discussion  of  the  stresses  in  disks. 

Mathematical  Treatment  of  Stresses  in  Disks.  Fig.  210  shows 
a  section  of  a  turbine  wheel  cut  out  (i)  by  two  radial  planes 

making  the  angle  0  with 
R+dR  each  other,  and  (2)  by  the 

cylindrical  surfaces  with 
radiuses  of  r  and  r  +  dr. 
The  two  other  bounding 
surfaces  are  the  sides  of 
the  disk.  The  thicknesses 
of  the  disk  are  t  at  the 
radius  r,  and  t  +  dt  at  the 
radius  r  +  dr. 

If  this  sector  is  rotated 
about  the  center  0  it 
develops  the  centrifugal 
force  (CF).  Acting  on  the 
surfaces  of  the  sector  are 

also  the  forces  R  and  R  +  dR  in  the  radial  direction  and  the  forces 
T,  T  in  tangential  directions.  The  two  tangential  forces  T,  T 
form  the  angle  180  —  6  degrees  with  each  other,  and  their  result- 
ant is  approximately  TO  when  6  is  a  small  angle.  We  have,  then, 

Forces  acting  outward  =  R  +  dR  +  CF. 
Forces  acting  inward     =  R  +  10. 

If  we  call  the  unit  stress  in  the  radial  direction  Sr  and  in  tan- 
gential direction  S<,  then  at  a  section  at  radius  r  (if  all  the  dimen- 
sions are  in  inches)  the  following  relations  result: 

R  =  r0tSr, 

R  -f  dR  =  (r  +  di)  (t  +  dt)  (Sr  +  </Sr)fl, 
T  = 

id  = 


FIG.  210.     Diagram  of  Disk  Stresses. 


STRESSES  IN  RINGS,  DRUMS,  AND  DISKS  323 

If  V  is  the  velocity  in  inches  per  second  and  w  is  the  weight 
of  a  cubic  inch  (the  specific  weight),  then  the  volume  of  the 
sector  is  very  nearly  tdidr,  and 

wV2 

CF  =  Wdr  —  • 
g 

For  equilibrium,  the  sum  of  the  forces  acting  outward  equals  the 
sum  of  those  acting  inward,  or 

R  +  dR  +  CF  =  R  +  T0,  or 


(r  +  dr)  (t  +  di)  (Sr  +  dSr)  0  +  =  r0tSr  +  tdrdSt. 

Dividing  through  by  0  and  neglecting  infinitesimals  of  the  second 
order,  we  have 

r(tdSr  +  Srdt)  +  tdr  (Sr  -  SO 


This  general  equation  is  not  suitable  for  calculations,  but  by 
assuming  conditions  of  uniform  strength  or  uniform  thickness 
the  form  can  be  considerably  simplified. 

Disk  of  Uniform  Strength.  If  we  assume,  then,  uniform  strength 
in  the  disk,  the  stresses  throughout  are  constant,  and  if  S'  is  the 
stress  at  any  point,  then 

S'  =  Sr  =  S*  =  constant  value 
and  therefore 

dS'  =  o,  and  substituting  these  values  in  equation  (34) 

TdtS'  +-tV2dr  =o, 
g 

dt    ,  wV2<fr 

T     ls^==0' 

—  +  -^-  X  -  -  -  =  o,  and  by  integrating, 
t       gS'  r 


*  K  is  a  constant  of  integration. 


324 


THE  STEAM  TURBINE 


Now  when  r  =  o,  t  =  t0,  and  K  =  —  log  t0,  so  that 


-;X+  (-logt.)  =o, 


w 


wV2* 

2gS" 


t    =  t0f 


(35) 


in  which  t  is  the  thickness  of  the  required  section  at  the  radius  r, 
t0  is  the  thickness  at  the  center,  and  e  is  the  base  of  Naperian  or 
natural  logarithms  which  is  equal  to  2.7183  and  Iog10  e  =  0.43429. 
All  the  symbols  in  these  equations  (including  2g  =  773  inches) 
are  in  inch  units. 

If  tj  is  the  minimum  thickness  of  the  disk,  then  equation  (35) 

can  be  written  wcw  -  v^) 

t=V      **s'     y  (36) 

where  Vj  is  the  peripheral  velocity  at  the  radius  corresponding 
to  tt  and  V  is  the  velocity  corresponding  to  t  as  before. 

If  the  disk  is  not  made  of  uniform  strength  throughout,  then 
Vt  is  the  velocity  where  the  portion  designed  for  uniform  strength 
begins. 

Equations  (35)  and  (36)  are  generally  used  by  the  designers 

of  impulse  turbines,  and  for 
the  conditions  of  average  prac- 
tice they  are  sufficiently  accu- 
rate. 

Design    of    the    Rim.     An 

FIG.  211.    Section  of  a  Turbine  Wheel.      enlarged  section  or  rim  is  usu- 
ally required  at  trie  circumfer- 
ence of  a  disk  for  the  attachment  of  the  blades.     Stresses  in  this 
section  require  careful  consideration. 

In  Fig.  2ii,  tj  is  the  smallest  thickness  of  the  disk  where  it 
joins  the  rim  (at  the  radius  rt)  and  t2  is  the  thickness  and  b2  the 

*  The  change  from  linear  to  angular  velocity  was  made  to  make  integration 
simpler. 


c.  of  G. 


STRESSES  IN  RINGS,  DRUMS,  AND   DISKS  325 

breadth  of  the  rim  of  which  the  center  of  gravity  is  at  the  radius 
r2.*  Blades  attached  to  the  rim  produce  by  the  centrifugal  force 
due  to  their  weight  the  stress  S2  in  pounds  per  square  inch. 
Besides  this  there  is  exerted  on  the  section  of  the  rim  the  stress 
due  to  the  centrifugal  force  of  its  own  weight  and  also  the  radial 
stress  (Sr)  in  the  disk  exerted  over  the  thickness  tr  The  expan- 
sion due  to  these  forces  acting  on  the  rim  must,  for  equilibrium, 
be  equal  to  the  expansion  of  the  section  of  the  disk  where  it  joins 
the  rim.  The  sum  of  the  radial  forces  Fr  acting  on  the  rim  per 
inch  of  length  may  be  stated  then  as 


in  which  w  is  the  weight  of  a  cubic  inch  of  the  material  of  the 
rim,  V2  is  the  velocity  at  the  radius  r2  in  inches  per  second. 

Radial  expansion  of  the  rim  (>}2)  is  expressed  by  the  following, 
form  if  a  is  the  area  of  the  rim  section  in  square  inches  and  E 
is  the  modulus  of  elasticity  in  pounds  per  square  inch,  J 


*  Because  the  contraction  of  the  cross-section  due  to  stresses  at  right  angles 
(page  320),  has  been  neglected  in  the  derivation,  equation  (35)  should  not  be  used 
for  values  of  allowable  unit  stress  less  than  15,000  pounds  per  square  inch,  as  it  gives 
thicknesses  at  the  center,  for  low  stresses,  which  are  sometimes  considerably  too 
large.  Practical  designers  who  are  required  to  use  unusually  low  stresses  for  disks 
will  find  a  suitable  discussion  in  Jude's  The  Theory  of  the  Steam  Turbine,  pages 
188  to  204.  wV  2 

f  Centrifugal  force  due  to  a  weight  of  a  cubic  inch  at  r2  is  -  —  ,  which  becomes 

Srz 

when  multiplied  by  the  area  of  the  rim  section. 

It  is  easily  shown  that  the  tensile  stress  in  a  thin  cylinder  is 

?-&-* 

a 

where  s  is  the  elongation  per  unit  of  length  and  a  is  the  area  of  the  section.     Then 
the  total  elongation  of  the  circumference  (^)  is  • 


and  the  radial  elongation  (>l)  is  i         p 

,_  *_*£ 

2?r        Ea 


326  THE  STEAM  TURBINE 

Since  the  radial  and  tangential  stresses  in  the  disk  have  been 
made  equal  in  the  original  assumptions,  the  unit  elongations  in 
every  direction  must  be  equal,  so  that  the  linear  expansion  in  the 
length  r  is 


where  k  is  the  coefficient  of  the  contraction  of  the  cross-section 
for  stresses  at  right  angles  (see  page  321). 

For  conditions  of  equilibrium  obviously  X2  =  At,  and  substi- 
tuting equation  (37)  in  (38)  and  equating  to  (39)  we  have 


Usually  the  percentage  error  from  writing  rt  for  r2  is  very 
small,  so  that  we  have  in  simpler  form, 


from  which  either  b2  or  t2  can  be  solved.  In  most  cases,  how- 
ever, t2  is  determined  by  the  blade  dimensions,  so  that  b2  is 
expressed  thus: 


(40) 


Vl2-  (l-k)Sr  Vt2    -.78,. 


In  this  equation  the  stresses  are  in  pounds  per  square  inch, 
Vl  and  g  are  in  inches  per  second,  tt,  t2,  rt,  and  b,  are  in  inches. 
Minimum  Thickness  of  the  Disk.  The  thickness  of  large  disks 
at  the  smallest  section  is  not  determined  by  the  allowable  stress 
but  by  the  requirements  for  safe  transportation  and  by  the  liability 
of  thin  disks  to  become  distorted  and  unstable  in  balance.  Disks 
about  5  feet  in  diameter  should  have  a  minimum  thickness  of 
from  .4  to  .6  inch,  depending  on  the  quality  of  the  material  and  the 


STRESSES  IN  RINGS,  DRUMS,  AND  DISKS  327 

speed  for  which  they  are  to  be  used;  and  for  disks  10  feet  in  diam- 
eter the  minimum  thickness  should  be  from  .7  to  1.25  inches.* 

The  breadth  of  the  rim  (b2)  calculated  by  equation  (40)  is 
the  maximum  value  allowable,  but  the  breadth  can  be  made,  of 
course,  less  than  that  calculated.  There  will  be  a  smaller  radial 
force  at  the  rim  of  the  disk  than  is  necessary  to  produce  the  uni- 
form radial  stress  Sn  and  the  disk  will  not  be  one  of  uniform 
strength.  The  stress  at  the  center  will  be  reduced  very  much 
less  than  that  at  the  smallest  section. 

If  now  equation  (40)  is  used  to  calculate  the  minimum  thickness 
tj,  with  an  assumed  value  for  b2  suitable  for  the  design,  negative 
values  may  be  obtained.  In  this  case  a  smaller  value  of  Sr 
must  be  used  in  the  calculation.  Limits  for  Sr  can  be  easily 
determined  by  putting  S2  =  o  in  (40);  then 

wV2 

ci 

which  is  the  tangential  stress  in  a  freely  rotating  ring  or  is  the 
usual  "fly-wheel"  formula  when  k=  o. 

Practical  Example.  Design  of  the  Rim  of  a  Disk  Wheel. 
A  disk  wheel  50  inches  in  diameter  is  to  be  designed  for  an  im- 
pulse turbine  to  operate  at  3000  r.p.m.  The  minimum  thickness 
(tt)  is  .4  inch,  and  nickel  steel  is  to  be  used  with  an  allowable 
stress  of  28,000  pounds  per  square  inch,  which  weighs  .28  pound 
per  cubic  inch  (w).  Approximately  the  radius  (rx)  at  the  inner 
edge  of  the  rim  is  25  inches,  so  that  V  is  7860  inches  per  second 
(about  450  miles  per  hour).  The  wheel  is  to  carry  two  rows 
of  blades,  so  that  the  thickness  of  the  rim  must  be  made  about 
3.5  inches.  The  weight  of  these  blades  is  equivalent  to  a  solid 
ring  of  steel  around  the  rim  .3  inch  thick,  f  The  weight  of  the 

*  Minimum  thickness  for  a  wheel  3  feet  in  diameter  is  about  .25  inch.  An 
approximate  rule  for  the  minimum  thickness  of  disks  is 

t  min  =  .008  d  to  .01  d, 
where  d  is  the  diameter  in  inches. 

f  Centrifugal  force  of  the  blades  on  a  wheel  is  probably  most  simply  determined 
by  this  method  of  calculating  from  a  drawing  showing  the  dimensions  of  the  blades 
the  thickness  of  a  solid  band  or  ring  of  the  same  weight. 


328  THE  STEAM  TURBINE 

blades  per  square  inch  of  the  rim  surface  is  .3  X  .28  =  .084 
pound,  and  the  stress  S2  per  square  inch  due  to  this  weight  is 
(take  g  =  386  inches  per  second) 

—  =  538  pounds  per  square  inch. 

25 

Substituting  these  values  in  (40), 

28,000  X  —  -  538 
b2  =    .28  X  (786o/'  -  X  25  -  1.5  inches. 

~  -  '7  ) 


The  thickness  of  the  section  at  the  center  (t0)  is  calculated  by 
(35),  using  the  same  allowable  stress  as  before  for  S': 

-  .28  (786o)2 

t  =  toe  772  x  28,000^ 

t0    =  te'*«    =  2.23t, 

t0  =  .4  X  2.23  =  .89  inch. 

The  expansion  of  the  radius  due  to  the  allowable  stresses  in  the 
disk  can  be  calculated  by  (39),  taking  E  =  30,000,000  pounds 
per  square  inch  and  k  =  .3, 

,  I    ;;  ^nr^1"  :  .  :•"  :   ::          ';. 

,       .7  X  28,000  X  25  ,  .     , 

L  =  -*  -  -  -  *  =  .016  inch, 
30,000,0000 

and  the  expansion  of  the  diameter  is  .032  inch. 

If  reaction  turbines  are  to  be  operated  at  higher  peripheral 
speeds  than  350  feet  per  second,  the  stresses  due  to  the  cen- 
trifugal forces  are  too  large  to  use  a  free  drum  construction,  so 
that  the  drum  must  be  strengthened  with  spokes  or  flat  disks. 
It  is  considered  better  practice,  however,  to  divide  a  drum  into 
short  sections,  and  calculate  each  section  by  the  method  explained 
here  for  disk  wheels  by  the  use  of  equations  (35)  to  (39).  The 
Allis-Chalmers  Company  uses  this  method  for  the  low-pressure 
stages  of  its  latest  designs  as  shown  in  Fig.  209,  although  the  per- 
ipheral speed  of  this  section  of  the  drum  is  usually  less  than  250 
feet  per  second. 


STRESSES  IN  RINGS,  DRUMS,  AND  DISKS  329 

Practical  Example.     Design  of  a  Wheel  Disk  without  a  Hole. 

Stresses  in  disks  are  difficult  because  the  areas  over  which  the 
forces  are  distributed  are  not  readily  determined;  besides,  the 
forces  are  not  uniformly  distributed  over  any  one  of  the  areas 
to  be*  considered.  The  stresses  in  a  disk  are  calculated  usually 
by  determining  the  force  acting  on  the  " boundary"  areas  of  a 
circular  sector  imagined  cut  out  of  the  disk.  Such  a  sector  is 
shown  in  Fig.  212.  The  radius  is  r  inches,  the  elementary  radial 
thickness  is  dr,  t  is  the  thickness  of  the  sector  (measured  parallel 


FIG.  212.     Forces  in  a  Sector  of  a  Wheel  Disk. 

to  the  axis  of  the  shaft),  and  0  is  the  angle  subtended  at  the 
center  by  this  sector.  The  centrifugal  forces  cause  tangential 
and  radial  stresses.  If  we  imagine  the  disk  made  up  of  a  series 
of  concentric  rings,  laid  side  by  side  and  touching,  the  tangential 
forces  tend  to  break  the  rings  in  the  line  of  the  tangent,  and  the 
purely  radial  forces,  on  the  other  hand,  will  tend,  as  it  were,  to 
break  out  pieces  which  would  be  carried  away  in  a  radial 
direction.  In  Fig.  212  the  tangential  and  radial  forces  are 
shown  more  simply  than  in  Fig.  210  in  the  directions  to  equili- 
brate the  centrifugal  force  CF.  In  other  words,  the  tangential 
and  radial  forces  shown  are  those  balancing  the  centrifugal 
forces. 


330  THE  STEAM  TURBINE 

An  actual  design  of  a  5o-inch  plain  disk  of  forged  steel  without 
a  hole  at  the  center  for  a  Riedler-Stumpf  turbine  is  shown  in 
Fig.  213,  and  the  following  paragraphs  show  how  the  calculations 
were  made.  Diameter  of  the  disk  (dj  is  46  inches  (measured 
inside  the  rim,  which  is  2  inches  wide).  Smallest  section  of  the 
disk  is  taken  as  .5  inch  (see  page  326).  Speed  is  4000  revolutions 
per  minute.  The  allowable  unit  stress  is  20,000  pounds  per  square 
inch,  and  the  disk  is  designed  for  uniform  strength.  Weight  of  the 
blades  is  .09  pound  per  inch  of  the  circumference,  producing  a 
centrifugal  force  of  1840  pounds  per  square  inch  at  the  smallest 
section  of  the  disk.  Now  it  has  been  shown*  that  in  a  flat  disk 
the  stress  at  the  edge  due  to  an  external  centrifugal  load  (like 
blades  and  shrouds)  is  superposable  by  simple  addition  to  the 
stresses  (both  radial  and  tangential)  in  the  disk  due  to  its  own 
rotation. 

The  rim  was  calculated  as  in  the  previous  example,  and  the 
thickness  (t2  in  Fig.  211)  was  determined  by  the  width  required 
for  the  blades.  From  the  inner  edge  of  the  rim  the  disk  was 
given  a  constant  thickness  of  .5  inch  till  the  tangential  stress 
alone  as  calculated  by  equation  (30)  exceeded  the  allowable 
limit,  f 

*  Jude,  The  Theory  oj  the  Steam  Turbine,  page  198. 

t  It  will  be  observed  that  these  approximations  are  very  much  on  the  safe  side 
because  of  the  effect  of  "  forces  at  right  angles  "  (see  page  320).  It  is  probable, 
however,  that  whatever  the  form  of  the  disk  (if  not  abnormally  irregular),  the 
stresses  at  the  center  are  slightly  higher  than  the  peripheral  stresses.  In  case  of 
undue  racing  due  to  the  failure  of  the  governing  apparatus  or  other  cause,  a  disk 
designed  for  uniform  strength  will  fly  to  pieces  from  the  center.  A  De  Laval 
wheel  without  the  usual  "safety  groove"  near  the  rim  when  tested  to  destruction 
broke  up  entirely  and  projected  large  pieces  through  a  cast-steel  casing  two  inches 
thick.  When,  however,  the  customary  groove  was  cut  just  inside  the  rim,  only 
pieces  of  the  rim  were  broken  off  when  an  excessive  speed  was  reached,  and  no 
external  damage  was  done. 

It  is  stated  by  Jude  that  the  metal  left  between  the  "safety  grooves"  of  a  De 
Laval  wheel  is  "only  sufficient  to  carry  the  traction  load  of  the  vanes."  From 
this  fact  the  minimum  thickness  of  De  Laval  disks  can  be  easily  calculated, 
as  it  is  generally  stated  that  the  factor  of  safety  at  the  groove  is  5,  and  the  section 
before  the  groove  is  cut  is  two-fifths  larger.  Allowable  unit  stress  is  probably 
taken  at  about  30,000  pounds. per  square  inch. 


STRESSES  IX  RINGS,  DRUMS,  AND  DISKS 


331 


For  some  distance  from  the  rim  toward  the  center  we  have  the 
case  of  a  flat  disk.  Now  in  a  disk  of  constant  thickness  without 
a  hole  at  the  center  both  the  radial  and  tangential  stresses  increase 
from  the  rim  toward  the  center.  At  the  outer  edge  of  such  a  disk 
the  radial  stress  is  only  that  due  to  the  cen- 
trifugal force  of  the  blades  and  rim,  while  the 
tangential  stress  is  of  considerable  magni- 
tude and  is  always  greater  than  the  radial 
stress,  except  at  the  center,  where  they  are 
equal.*  Because  both  radial  and  tangential 
stresses  at  every  section  are  approximately 
increased  1840  pounds  per  square  inch  by  the 
centrifugal  force  due  to  the  blades,  the  net 
allowable  stress  is  18,160  pounds  per  square 
inch.  The  point  where  the  "increasing"  sec- 
tion begins  is  determined  then  by  the  following 
calculations  —  substituting  in  equation  (30): 


18,160  X  139,100 

.28  X  (4000 )2 
dt  =  23.8  inches,  or  rt 


=  565, 

=  11.9  inches. 


Beyond  this  point  toward  the  center  of  the 
disk  the  section  has  been  made  of  uniform 
strength  as  calculated  by  equation  (35).  The 
calculation  of  the  thickness  where  the  diameter 
is  10  inches  (r  =  5  inches)  is  given  by  equa-  _ 
tion  (36),  FIG.  213. 


.28X21.2(10)* 

j  723  X  18,160 


Design  of  a 
Wheel  Disk  without 
a  Hole  at  the  Center. 


in  which  t^  =  .5  inch  and  e  =  2.7183,  then, 

t  =  V45 

t  =  1.56  tt  =  .78  inch. 

In  the  same  way  the  thickness  can  be  calculated  for  enough  points 
to  determine  the  profile.      The  section  shown  in  Fig.  213  is  a 

*  Jude,  The  Theory  of  the  Steam  Turbine,  page  200. 


332  THE  STEAM  TURBINE 

typical  "flat"  disk.  To  facilitate  the  forging  of  such  disks  the 
profile  is  not  made  exactly  as  calculated  but  is  gradually  tapering 
from  the  smallest  section  to  the  center.  The  increase  in  the 
thickness  from  the  rim  to  the  center  is  very  small  compared  with 
many  designs,  approximating  a  "concavo-convex"  form  (Fig. 
214).  It  is  argued  by  some  designers  that  the  treatment  of  the 
"concavo-convex"  forms  is  entirely  wrong  and  that  most  prob- 
ably it  is  not  possible  for  the  stresses  all  along  the  central  plane 
to  be  either  equal  to  or  less  than  those  at  the  rim  by  merely  satis- 
fying equation  (35),  and  that  the  metal  in  the  bulging  part  of  such 
disks  has  little  influence  in  modifying  the  stresses  in  the  central 
plane.  Some  of  the  best  authorities  agree  that  it  seems  reason- 
able that  whatever  the  form  of  the  profile  of  a  disk  the  "stresses 


j^r-     -] 
FIG.  214.     Typical  Solid  Disk  without  a  Hole  at  the  Center. 

in  and  about  the  central  plane  do  not  differ  greatly  from  those 
in  a  flat  disk  running  at  the  same  speed."* 

A  typical  solid  disk  without  a  hole  for  bolts  or  for  the  passage  of 
a  shaft  through  it  is  shown  in  Fig.  214.  It  was  designed  for  a 
very  much  higher  speed  than  the  one  in  Fig.  213,  so  that  it  has 
a  bulging  from  near  the  center.  This  design  shows  an  ingenious 
method  for  the  attachment  of  the  body  of  the  disk  to  the  shaft. 
It  will  be  observed  that  the  disk  is  made  with  a  very  small  sec- 
tion near  the  rim,  so  that  the  stress  there  far  exceeds  that  any- 
where else.  If  the  wheel  breaks  it  will  rupture  first  at  this 
smallest  section  and  the  rim  and  blades  will  be  torn  off.  When 
these  parts  are  gone  the  centrifugal  force  will  be  so  much  reduced 
on  the  part  of  the  wheel  remaining  that  there  can  be  still  a  very 

*  Jude,  The  Theory  of  the  Steam  Turbine,  page  204. 


STRESSES  IN  RINGS,  DRUMS,  AND  DISKS  333 

great  increase  in  speed  without  further  damage.  This  disk  is 
designed  for  a  factor  of  safety  of  about  five  at  the  smallest  section, 
and  about  seven  at  every  other  section. 

In  designing  a  disk  for  high  speeds,  obviously  a  section  that 
gives  approximately  uniform  strength  from  the  rim  to  the  center 
is  desirable.  Experience  in  such  calculations  has  shown  that  a 
disk  of  the  shape  shown  in  Fig.  211  fulfills  approximately  these 
conditions.*  This  disk  was  designed  for  a  speed  of  20,000 
revolutions  per  minute.  There  is  a  centrifugal  force  of  about 
.2  pound  per  inch  on  the  outside  of  the  rim,  due  to  the  weight  of 
the  blades  which  are  of  the  irregular  shape  shown  in  Fig.  64. 

Disks  with  Holes  in  the  Center.  Up  to  this  point  in  the  dis- 
cussion of  stresses  in  disks  only  designs  similar  to  Figs.  213  and 
214,  without  a  hole,  have  been  considered.  When,  however,  a 
hole  is  made  near  the  center  of  a  disk  the  stresses  are  greatly 
increased.  There  are  no  very  reliable  methods  for  determining 
the  stresses  in  disks  of  arbitrary  shapes  with  central  holes. 
According  to  De  Laval,  any  methods  for  "taking  into  account 
the  hub  influences  in  the  calculation  are  only  rough  approxi- 
mations" to  the  actual  conditions.  It  can  be  shown  theoretically 
that  a  mere  pin-hole  at  the  center  of  a  disk  makes  the  tan- 
gential stress  St  at  the  hole  twice  that  in  a  disk  without  a  hole. 
Indeed  a  small  flaw  near  the  center  of  a  disk  may  seriously 
affect  the  magnitude  of  the  stresses.  For  this  reason,  steel  ingots 
with  any  traces  of  " piping"  must  not  be  used  for  forged  disks 
to  be  operated  at  high  speeds.  For  thick  disks  of  the  typical 
De  Laval  shape  when  perforated,  the  exact  solution  is  appar- 
ently indeterminate.  Methods  of  calculation  for  such  irregular 
sections  have  been  proposed  which  depend  on  the  determination 
of  the  mean  stresses  of  the  whole  section.  Results  from  such 
methods  are,  however,  of  no  value  at  all,  as  it  is  known  that  the 
maximum  stresses  are  often  twice  the  calculated  mean  stress. 

*  Besides  blow-holes  and  piping  in  ingots  for  drop  forgings,  most  makers  put 
holes  into  the  disks  for  the  attachment  of  tools  for  removing  the  disks  from  the 
shafts,  and  for  balancing  weights.  Very  few  disks  are  made  that  do  not  have  some 
holes. 


334 


THE   STEAM  TURBINE 


The  fact  remains,  however,  that  disks  for  turbines  are  very 
commonly  made  with  holes  in  the  center  for  the  shaft,  and  other 
holes  besides  are  often  made  for  the  attachment  of  tools  for  forcing 
the  disk  from  the  shaft  when  the  wheel  is  to  be  removed.  Stress 
distribution  near  a  central  hole  of  a  nearly  flat  disk  can  be  approx- 
imately calculated  if  a  disk  of  comparatively  smaller  diameter  is 
imagined  cut  from  its  center,  and  this  small  disk  is  then  assumed 
to  be  of  constant  thickness  and  subjected  to  a  radial  stress  at  its 

rim  equal  to  the  uniform 
stress  in  the  large  disk  if  it 
had  no  hole.  The  stresses 
in  this  small  disk  with  the 

g  soixx)- 1^  hole  can  be  calculated  with 

some  degree  of  accuracy 
from  equation  (34)  by  put- 
ting dt  =  o,  since  t  has  been 
assumed  constant  in  this 
small  disk.*  Tangential 
stresses  calculated  in  this 
way  for  a  disk  10  inches  in 
diameter  with  a  hole  i  inch 
in  diameter  are  shown  in 
Fig.  215.  Radial  stress  is, 

of  course,  zero  at  the  center  so  that  it  is  not  important.  The 
large  disk  (Fig.  213)  was  designed  to  make  the  combined  unit 
stress  in  the  section  20,000  pounds  per  square  inch,  and  it  is 
assumed,  therefore,  that  the  radial  stress  on  the  outside  of  the 

*  Simplified  formulas  for  a  disk  of  constant  thickness  are  given  by 
Eyerman,  Die  Dampf turbine,  pages  88-90  ;   Stodola,  Die  Dampfturbinen,  pages 
160-161. 

The  algebraic  work  involved  in  obtaining  equations  suitable  for  calculations  is 
laborious  and  complicated.  Because  these  equations  are  not  used  directly  for  other 
calculations  they  are  not  given  here.  This  chapter  on  stresses  is  not  intended  to  be 
an  exhaustive  treatment,  mathematically,  and  the  practical  designer  wishing  to  use 
the  minimum  factors  of  safety  should  carefully  study  the  graphical  solutions  given 
by  Stodola;  but  he  should  remember  that  these  methods  referred  to  are  only 
approximations  and  in  a  great  measure  are  justified  only  because  they  have  stood 
the  test  when  applied  in  practice. 


FIG.  215.     Variation  of  Stress  in  a  Disk 
caused  by  a  Hole  at  the  Center. 


STRESSES  IN  RINGS,  DRUMS,  AND  DISKS  335 

small  disk  has  this  value.  The  curve  shows  that  the  maximum 
stress  at  the  hole  is  40,000  pounds  per  square  inch  and  that  the 
stress  is  rapidly  reduced  as  the  distance  from  the  edge  of  the  hole 
increases  till  it  reaches  the  constant  value  of  20,000  pounds  per 
square  inch,  for  which  the  wheel  disk  was  designed.  It  should  be 
observed,  therefore,  that  the  stress  at  the  edge  of  a  hole  at  the 
center  of  a  disk  is  twice  that  at  some  distance  away  from  the  hole.* 
It  should  be  carefully  noted,  however,  that  this  discussion  applies 
only  to  holes  at  the  center  of  a  disk.  Holes  near  the  rim  such 
as  are  often  made  for  balancing  the  disk  or  as  a  safety  device 
so  that  the  rim  will  break  first  in  case  of  excessive  speed, 
would  be  allowed  for  in  practice  merely  by  the  reduction  of  the 
section. 

It  is,  however,  a  good  practice  to  make  the  section  at  the  hub 
of  a  disk  with  a  hole  at  the  center  of  sufficient  size  to  withstand 
the  greatest  stress  that  may  come  to  bear  at  the  normal  speed. 
Fig.  216  shows  how  the  disk  in  Fig.  214  should  be  modified  that 
it  may  be  put  on  a  4- inch  shaft.  The  thickness  (z)  of  the  hub 
will  be  determined  in  the  usual  way  as  discussed  in  books  on 
machine  design.  Only  its  length  (t0)  concerns  this  discussion. 
Eyermanj  and  Stodola  give  elaborate  graphic  methods  for  this 
determination,  but  they  will  not  be  taken  up  here,  as  they  are  of 
no  general  interest.  For  most  practical  purposes  it  is  satisfactory 
to  make  use  of  the  results  shown  in  Fig.  215  and  make  the  length 
of  the  boss  (t0)  twice  the  thickness  at  the  same  section  for  a 
disk  without  a  hole.  Instead  of  reducing  the  section  abruptly 
in  proportion  to  the  reduction  in  the  stress  the  use  of  a  fillet 
(see  curve  ab  in  Fig.  216)  of  very  gentle  curvature  gives  by  far 
the  best  construction.! 

Because  the  distribution  of  stress  is  changed  when  a  hole  is 
made  in  the  center  of  a  flat  disk,  the  section  where  the  radial 

*  This  has  been  shown  by  a  mathematical  demonstration  and  the  development 
of  suitable  formulas  by  Grubler,  Zeit.  Verein  deutscher  Ingenieure,  1897,  page 
860  ;  Kirsch,  Zeit.  Verein  deutscher  Ingenieure,  1897,  page  798. 

f  W.  Eyerman,  Die  Dampfiurbine,  pages  86-98. 

J  Stodola,  Die  Dampfturbinen,  3rd  edition,  page  164. 


336 


THE  STEAM  TURBINE 


stress  equals  the  allowable  limit  must  be  calculated  by  a  different 
method  from  that  used  for  the  disk  without  a  hole  at  the  center. 

To  determine  from  the  general  theoretical  equations  for  the 
stresses  in  disks  the  diameter  where  the  radial  stress  in  a  flat 
disk  with  a  hole  at  the  center  has  a  definite 
value  is  very  laborious  and  almost  impracti- 
cable; but  the  following  approximate  and  more 
or  less  empirical  formula  for  the  radial  stress 
in  a  disk  of  uniform  thickness  with  a  hole  in 
the  center  can  be  used  conveniently.  It  is 
practically  the  same  as  that  given  by  Cree  and 
Jude  *  except  that  it  has  been  simplified  by 
grouping  constants  and  changing  the  units  to 
correspond  with  those  used  in  the  other  equa- 
tions in  this  chapter. 

If  Sr  is  the  radial  stress  in  the  disk  in  pounds 
per  square  inch  at  any  diameter  dt  inches,  V  is 
the  velocity  at  the  periphery  of  the  disk  in  feet 
per  second,  D  is  the  diameter  of  the  disk  in 
inches,  and  d  is  the  diameter  in  inches  of  a  hole 
at  the  center,  then 


Now  if  this  equation  is  to  be  solved  to  deter-     -s|~ 
mine  dly  it  can  be  written  i 


d,4  +(^|r  Sr-0'-d')d1'- 

and  putting  B  =  fi~^  Sr  -  D2  -  d2J 


D2d2  =  O,  (42)     FIG.  216.     Design 
of  a  Wheel  Disk 
with   a  Hole  at 
and  the  Center. 


C  =  D2d2,  then 


(43) 


*  Jude,  The  Theory  of  the  Steam  Turbine,  page  204. 
\ 


STRESSES  IN  RINGS,  DRUMS,  AND  DISKS  337 

This  last  equation  is  easily  solved  after  obtaining  the  values  of 
B  and  C  from  the  dimensions  of  the  disk  and  the  allowable  unit 
stress. 

There  are  two  values  of  d,  because  the  radial  stress  increases 
to  a  maximum  value  and  then  decreases  to  zero  at  the  edge  of 
the  hole.  The  larger  value  of  dt  is  always  taken  to  determine 
the  design  because  between  the  two  values  of  dl  the  radial  stress 
has  its  maximum  value. 

In  the  design  for  this  example  Sr  =  18,160  pounds  per  square 
inch,  D  =  46  inches,  d  =  4  inches,  and  V  =  500  feet  per  second. 
The  value  of  B  is  then  1358,  C  is  33,856,  and  dt  is  calculated  to  be 
36.5  inches. 

The  section  from  the  36.5  inch  diameter  inward  toward  the 
center  is  made  of  uniform  strength  and  is  calculated  by  the  use 
of  equation  (36)  in  the  same  way  as  in  the  preceding  examples. 

Permissible  Stresses  and  Suitable  Materials.  It  is  considered 
safe  generally  to  use  ordinary  forged  or  rolled  steel,  for  velocities 
not  exceeding  600  feet  per  second;  and  for  lower  speeds  than 
this  limit  wrought  iron  can  even  be  used  if  it  is  of  exceptionally 
good  quality.  For  speeds  from  600  to  1000  feet  per  second 
crucible  cast  steel  can  be  used. 

Nickel  steel  is  recommended  for  turbine  disks  by  the  Krupp 
Company  of  Essen,  Germany.  This  nickel  steel  has  an  ulti- 
mate tensile  strength,  of  125,000  pounds  per  square  inch  and 
12  per  cent,  elongation  before  rupture.  The  elastic  limit  is 
about  95,000  pounds  per  square  inch.  It  is  stated  by  the 
Krupp  Company  that  they  will  produce  a  nickel  steel  of  still 
higher  tensile  strength  but  only  about  6  per  cent,  elongation. 
With  some  small  forged  pieces  of  this  material  an  ultimate 
tensile  strength  of  285,000  pounds  per  square  inch  has  been 
observed,  with  an  elastic  limit  of  nearly  225,000  pounds  per 
square  inch.  All  De  Laval  turbine  wheels  used  in  America  are 
made  in  Sweden  of  forged  nickel  steel,  which  is  rather  high  in 
carbon. 

Allowable  working  stresses  must,  of  course,  be  left  to  the 
judgment  of  the  designers.  An  engineer  of  the  Krupp  Com- 


338  THE   STEAM  TURBINE 

pany  states  that  stresses  in  the  same  direction  may  be  allowed 
in  turbine  disks  as  high  as  one-third  of  the  elastic  limit. 

Since  the  centrifugal  force  and  therefore  also  the  unit  stress  is 
proportional  to  the  square  of  the  velocity,  if  a  factor  of  safety 
of  4  is  allowed,  the  breaking  speed  of  the  wheel  will  be  twice  the 
normal  speed,  and  the  elastic  limit  of  the  material  is  only  about 
1.5  times  the  normal  speed. 

.Excessive  stresses  at  a  hole  are  " dissipated"  very  materially 
if  a  dangerous  stress  is  reached  at  the  edge  of  the  hole.  Before 
rupture  can  occur  there  will  be  an  excessive  elongation  of  the 
material  as  soon  as  the  elastic  limit  is  reached  at  the  highly 
stressed  section. 

CRITICAL   SPEEDS   OF  LOADED   SHAFTS. 

With  the  high  speeds  at  which  steam  turbines  are  operated 
the  centrifugal  forces  due  to  even  a  small  eccentricity  of  the 
rotating  masses  produce  vibrations,  excessive  stresses,  and 
"springing"  of  shafts.  As  the  result  of  the  eccentric  forces  the 
shaft  is  bent  farther  out  of  line,  so  that  the  centrifugal  forces 
and  the  amount  of  the  eccentricity  are  increased  until  the  stress 
set  up  in  the  shaft  by  the  bending  produces  a  force  equal  to  the 
centrifugal  force,  and  the  center  of  gravity  and  "  center  of  work" 
coincide.  If  W  is  the  weight  of  the  rotating  mass  in  pounds,  e  is 
the  "original"  eccentricity  of  the  shaft  in  inches,  x  is  the  eccen- 
tricity in  inches  at  N  revolutions  per  minute,  P  is  the  force 
applied  to  the  shaft  at  the  point  of  attachment  of  the  disk  which 
will  bend  the  shaft  i  inch,  within  the -elastic  limit,  C.  F.  is  the 
centrifugal  force  of  the  rotating  mass,  and  k  is  a  constant,  then 

c  F  =  kWN*. 

gx 

The  bending  of  the  shaft  at  this  speed  is  x  —  e,  so  that 

kWN2 

(x  -  e)P  =  (44) 

gx 

and  e  N 


STRESSES  IX  RINGS,  DRUMS,  AND  DISKS  339 

The  increased  eccentricity  due  to  rotation  is  therefore  propor- 
tional to  the  original  eccentricity  of  the  shaft  and  increases  with 
increasing  values  of  u  and  hence  also  of  N.  When 

kWu2 


x  becomes  oo  ,  that  is,  the  deflection  becomes  exceedingly  large, 
unless  prevented,  and  would  break  the  shaft. 

It  has  been  shown  by  Cree  *  that  the  critical  speed  Nc  (r.p.m.) 
of  a  shaft  with  some  flexibility  in  the  bearings  carrying  a  con- 
centrated load  of  W  pounds  is 


6.94r2  .  /El 

-' 


(45) 


where  E  is  the  modulus  of  elasticity  in  pounds  per  square  inch, 
r  is  the  radius  of  the  shaft  in  inches,  1  is  its  length  (between  two 
bearings)  in  feet,  and  a  and  b  are  the  distances  from  the  load  to 
the  bearings,  in  feet.  This  formula  f  is  to  be  used  for  only  a 
single  concentrated  load  like  the  single  wheel  of  a  De  Laval 
turbine.  When  there  are  a  number  of  wheels  with  possibly  also 
a  revolving  field  of  a  generator  on  the  same  shaft,  the  problem 
becomes  very  complicated  if  the  loads  are  considered  separately. 
Experience  with  such  calculations  has  shown  that  for  the  cases 
occurring  in  practice  $  the  critical  speed  can  be  determined  by  the 
following  simple  equation  derived  for  the  case  of  uniform  loading: 

Nc=  i55,ooo  r'Y^p'  (46) 

where  W  is  the  sum  of  the  several  loads  on  the  shaft  and  the 
other  symbols  are  used  as  before. 

*  Proc.  Physical  Society  (London),  vol.  XIX. 

t  In  this  formula  the  weight  of  the  shaft  is  not  taken  into  account.  The 
influence  of  the  weight  of  the  shaft  on  the  critical  speed  can  be  easily  cal- 
culated, but  in  practical  cases  it  may  be  neglected  without  appreciable  error. 

J  This  applies  particularly  to  the  cases  of  Rateau,  Parsons,  and  Curtis 
turbines  and  turbine-driven  generators  and  pumps. 


CHAPTER  XVI. 
GAS  TURBINES. 

THE  development  of  the  gas  turbine,  which  should  combine 
the  high  thermal  efficiency  of  an  internal  combustion  engine  with 
the  mechanical  simplicity  of  the  steam  turbine,  has  occupied  the 
attention  of  a  number  of  able  engineers  from  time  to  time  but 
without  unqualified  success.  Because  of  the  severe  conditions 
due  to  the  very  high  temperatures  of  the  gases  after  combustion, 
there  are  many  difficulties  in  construction  which  in  a  large 
measure  offset  the  otherwise  simple  mechanical  construction. 

It  may  well  be  said  that  the  designer  of  gas  turbines  is  between 
"the  two  horns  of  a  dilemma."  If  he  tries  to  utilize  the  gases 
at  the  temperatures  resulting  from  expansion  in  a  single  normal 
nozzle,  the  nozzles  and  blades  will  deteriorate  very  rapidly,  and 
for  the  best  efficiency  the  speed  of  rotation  of  the  turbine  must 
be  made  too  high  for  utilization  for  general  power  purposes  with- 
out the  application  of  reducing  gears;  and,  if  on  the  other  hand, 
he  cools  the  gases  by  the  injection  of  water  or  excess  air  into  the 
combustion  chamber  to  make  the  temperature  of  the  gases  suitable 
for  the  materials  available  for  machine  construction,  the  high 
thermal  efficiency  stated  by  the  simplest  laws  of  thermodynam- 
ics* is,  of  course,  not  attained. 

Since  the  gas  turbine  is  certainly  not  yet  out  of  the  experi- 
mental stage,  although  there  are  commercial  applications,  it  is 
not  out  of  place  to  give  some  space  to  its  history. 

Probably  the  oldest  form  of  gas  turbine  is  the  ancient  propeller 

*  The  thermodynamic  efficiency  of  a  heat  engine  is  expressed  by where 

1 1 

TI  is  the  initial  and  r2is  the  final  temperature  of  the  cycle.  By  lowering  the  value 
of  Tv  the  efficiency  is  reduced  in  much  greater  proportion  than  the  reduction  in  the 
temperature. 

340 


GAS  TURBINES 


341 


mechanism,  known  as  a  "smoke-jack,"  which  was  used  for 
operating  the  turnspit*  of  large  open  fireplaces.  An  illustra- 
tion of  this  "smoke-jack"  is  shown 
in  Fig,  218,  which  is  a  copy  of  an 
old  drawing  published  in  Bishop 
Wilkin's  Mathematical  Magic  in  1680. 
A  similar  apparatus  is  described  by 
Cardan  about  1550.  This  mechanism 
was  placed  in  the  chimney  and  was 
driven  around  by  the  ascending  cur- 
rent of  hot  gases  from  the  fire.  Its 
motion  was  transmitted  by  gearing 
and  belting  to  the  spit  on  which  the 
joint  of  meat  was  carried  in  front  of 

.        -.  ..    .  .     ..  FIG.  218.     A  Chimney  Turnspit 

the  fire.     The  power  of  this     smoke-  or  "Smoke-Jack  " 

jack"  can  only  be  estimated  by  the 

work  of  the  turnspit  dog  which  it  replaced.     It  must,  therefore, 
be  rated  at  least  one  "  dog-power. " 

The  earliest  attempt  to  construct  a  gas  turbine  on  scientific 
principles  was  probably  made  by  Stoltze  of  Charlottenburg,  who 
received  a  patent  for  what  he  called  a  " hot-air"  turbine  in  1873. 
This  apparatus  consisted  of  two  turbines  on  one  shaft,  one  acting 
as  an  air  compressor  and  the  other  as  a  power  turbine.  The 
function  of  one  of  these  turbines  was  to  draw  in  and  compress 
the  air  to  about  40  pounds  per  square  inch  absolute.  Part  of 
this  compressed  air  was  then  passed  through  a  combustion 
chamber  or  furnace,  where  it  supplied  the  oxygen  required  for 
the  combustion  of  the  gas  or  oil  fuel.  Another  part  went 
through  a  heating  chamber  and  was  later  mixed  with  the  gases 
of  combustion  from  the  furnace.  The  mixture  of  gas  and 
air  was  then  expanded  in  the  second"  turbine.  The  useful 
power  developed  by  such  a  turbine  is  the  difference  between 
that  developed  by  the  gas  turbine  and  that  required  to  drive 

*  Turnspit  is  the  name  usually  applied  to  the  dog  which  was  used  to  turn,  by 
means  of  a  suitable  mechanical  contrivance,  a  spit  or  long  iron  bar,  pointed  at  one 
end,  used  to  hang  up  meat  to  be  roasted. 


342 


THE  STEAM  TURBINE 


the  turbine-compressor.  A  turbine  designed  to  develop  200 
horsepower  has  been  constructed  on  this  plan,  but  it  has  not 
been  commercially  developed.  It  is  very  doubtful,  if  all  other 
difficulties  were  overcome,  whether  this  method  of  air  injection 
could  give  nearly  as  good  economy  as  water  injection.  (See 
page  344.) 

Some  attention  has  been  given  to  the  development  of  the 
explosion  gas  turbine,  of  which  a  very  simple  form  is  shown  in 
Fig.  219.  It  consists  of  a  combustion  chamber  E,  of  which  one 
end  is  closed  by  a  large  valve  A  opening  inward,  admitting  air 
through  the  parts  B,  B  and  fuel  through  tubes  F,  F  opening  into 
the  valve  seat.  The  mixture  of  gas  and  air  is  ignited  by  electric 
sparks  at  I,  and  the  products  of  combustion  are  discharged  from 


FIG.  219.     A  Simple  Explosion  Gas  Turbine. 

the  chamber  through  a  small  opening  J  leading  into  the  nozzle  N, 
where  air,  as  shown  by  the  arrows,  is  mixed  with  the  gases  to 
reduce  their  temperature  before  they  reach  the  blades  of  the 
turbine  wheel  W  opposite  the  nozzle. 

It  is  a  well-established  fact  that  when  a  mixture  of  gas  and  air 
is  exploded  there  is  first  a  sudden  expansion  and  then,  because 
of  the  combination  of  the  hydrogen  in  the  burned  gases  with 
the  oxygen  in  the  excess  air  to  form  water,  a  vacuum  is  produced. 
This  phenomenon  "is  applied  in  this  apparatus  to  operate  the 
valve  A,  which  by  the  formation  of  a  vacuum  is  drawn  inward 
to  admit  another  charge  of  gas  and  air.  It  is  stated  that  in 
such  a  turbine  the  explosions  will  occur  very  rapidly  —  from  3500 
to  5000  per  minute  —  so  that  there  is  a  practically  continuous 
discharge  upon  the  wheel.  The  efficiency  of  an  explosion  motor 
of  this  kind  is  very  low  because  of  the  lack  of  compression;  but 


GAS  TURBINES 


343 


its  efficient  development  does  not  seem  to  be  impossible.  If  in 
some  way  efficient  combustion  by  explosion  can  be  segured 
without  compression,  then  a  most  economical  power  development 
could  be  attained  with  an  explosion  combustion  chamber  with 
the  fuel  and  air  valves  operated  automatically  "by  vacuum"  and 
the  injection  of  probably  comparatively  large  quantities  of 
water  after  combustion.  Such  an  apparatus  would  be  simple 


Oil  Pump 

FIG.  220.     Section  of  a  Zoelly  Explosion  Gas  Turbine. 

indeed  compared,  on  the  one  hand,  with  the  complicated  com- 
bination of  the  steam  toiler  with  external  firing  and  the  steam 
turbine,  or,  on  the  other  hand,  with  the  complex  reciprocating 
gas  engine.  Fig.  220  illustrates  a  Zoelly  explosion  gas  turbine. 
It  consists  essentially  of  an  explosion  chamber  C,  a  turbine 
wheel  W,  water  and  oil  pumps,  and  an  air  compressor.  The 
pumps  and  compressor  are  of  the  reciprocating  type  and  are 
driven  by  the  main  shaft  by  means  of  the  worm  gears  Ax  and 
Bj.  The  valves  regulating  water,  oil,  and  air  admission  and 
the  ignition  device  TI,  are  operated  by  the  gases,  and  steams  are 


344  THE  STEAM  TURBINE 

expanded  in  the  nozzle  N  and  impinge  upon  a  turbine  wheel  W 
of  the,  De  Laval  type.  Some  of  the  heat  remaining  in  the  exhaust 
gases  is  absorbed  by  water  coils  R  which  serve  to  heat  the  injec- 
tion water.  In  the  operation  of  this  apparatus,  air  is  admitted 
first  into  the  explosion  chamber  and  then  the  oil,  as  the  air  is 
supposed  to  act  as  a  shield  against  back-firing.  After  the  charge 
has  been  exploded  and  the  maximum  pressure  has  been  reached, 
the  cooling  water  is  injected. 

The  more  successful  gas  turbines,  however,  are  those  operating 
by  combustion  at  constant  pressure.  In  this  type  the  air  and 
fuel  (oil  or  gas)  are  delivered  under  pressure  to  a  suitable  com- 


Water 


FIG.  221.     Diagrammatic  Illustration  of  the  Combustion  Chamber  and  Steam 
Coils  of  a  Modern  Gas  Turbine. 

bustion  chamber  A  in  Fig.  221  which  is  maintained  at  a  red 
heat,  so  that  the  combustion  is  continuous.  The  products  of 
combustion  are  usually  cooled  by  water  which  is  injected  into  the 
nozzle  as  in  the  explosion  type.  The  heat  energy  in  the  burned 
gases  is  converted  into  velocity  in  an  expanding  nozzle  N  and  are 
discharged  at  a  high  velocity  upon  the  blades  R  of  the  turbine 
wheel.  Designers  of  this  type  of  gas  turbines  have  generally 
assumed  that  nozzles  and  wheels  of  the  De  Laval  type  are  most 
suitable,  arid  their  energies  are  devoted  at  present  to  the  pro- 
duction of  a  suitable  combustion  apparatus  and  a  high  efficiency 
rotary  compressor.  Fig.  222  shows  a  typical  small  gas  turbine 
set  up  for  a  brake  test. 


GAS  TURBINES 


345 


FIG.  222.     A  Gas  Turbine  set  up  for  a  Brake  Test. 


346 


THE  STEAM  TURBINE 


In  practice  the  combustion  chamber  is  lined  with  carbo- 
rundum, and  to  allow  for  expansion  the  carborundum  is  backed 
with  sheets  of  asbestos  to  provide  a  soft  and  elastic  packing. 
Exhaust  gases  are  usually  discharged  over  a  coil  boiler  L,  and 
the  steam  which  is  produced  is  also  delivered  upon  the  turbine 


FIG.  223.      Arinengaud  and  Lemale's  Gas  Turbine. 

wheel  by  a  separate  nozzle  M.  When  the  turbine  is  in  operation 
the  lining  becomes  sufficiently  hot  to  ignite  the  fuel  as  it  is 
forced  into  the  chamber. 

A  gas  turbine  of  this  latter  type  designed  by  Arinengaud  and 
Lemale  of  Paris  is  illustrated  in  Fig.  223.  It  is  a  machine 
developing  300  net  horsepower  at  4000  revolutions  per  minute. 


GAS  TURBINES 


347 


A  Rateau  turbine-compressor  shown  direct  connected  to  the  gas 
turbine  in  Fig.  224  has  been  specially  designed  and  built  by 
Brown,  Boveri  &  Co.  of  Baden,  Switzerland,  for  use  with  this 
turbine.  The  compressor  gives  a  mechanical  efficiency  as  high 
as  65  to  70  per  cent,  and  delivers  i  cubic  foot  of  air  per  second 


FIG.  224.      Arinengaud  and  Lemale's  Gas  Turbine  Direct  Connected  to  a 
Rateau  Turbine-Compressor. 


at  a  pressure  of  from  6  to  7  atmospheres.  Compressed  air  is 
used  for  starting,  and  a  simple  ignition  device  is  used  for  firing 
the  charge  till  the  combustion  chamber  becomes  sufficiently 
heated.  M.  Barbezat,  who  has  now  charge  of  the  development 
of  this  turbine,  states  that  the  total  efficiency  is  not  as  high  as 
that  of  reciprocating  gas  engines;  but  no  data  are  given. 


34$  THE  STEAM  TURBINE 

Gas  turbines  have  been  applied  practically  for  the  propulsion 
of  submarine  torpedoes.  Formerly  some  types  of  torpedoes 
received  their  motive  power  from  a  rotary  motor  like  a  turbine 
wheel,  driven  by  compressed  air.  Recently  gas  turbines  have 
been  installed  with  an  obvious  gain  in  power  and  saving  in  weight. 
These  gas  turbines  develop  120  horsepower  at  1000  revolutions 
per  minute.  The  expansion  ratio  of  the  nozzles  is  8.4  and  the 
weight  per  horsepower,  without  the  compressor,  is  1.3  pounds. 

It  is  obvious,  then,  that  great  progress  has  been  made  recently 
in  the  development  of  the  gas  turbine;  and  when  the  ratio  of 
progress  is  compared  with  the  time  required  to  bring  the  recip- 
rocating gas  engine  to  its  present  state  of  development,  there  is 
reason  for  hoping  for  greater  accomplishments  in  the  near  future. 
The  gas  turbine  question  includes,  however,  a  number  of  un- 
solved problems;  but,  on  the  other  hand,  the  sources  available 
for  their  solution  are  numerous.  The  development  of  these 
machines  will  permit  the  utilization  for  power  of  mixtures  of  air 
with  coal  gas,  petroleum,  or  alcohol;  and  it  will  also  make  possible 
a  combination  of  the  explosion  motor  and  the  steam  turbine  for 
many  purposes. 

The  problem  is  laid  plainly  before  the  physicist,  the  engineer, 
and  the  machinist,  and  to  bring  about  a  satisfactory  solution  will 
doubtless  require  all  their  combined  resources. 

Questions  of  Theory.  The  success  of  the  steam  turbine 
naturally  directed  the  attention  of  engineers  to  the  possibilities 
of  the  gas  turbine  with  the  expectation  of  combining  the  high 
thermal  efficiency  of  the  gas  engine  with  the  constructive  advan- 
tages of  the  steam  turbine. 

As  explained  in  the  preceding  pages  a  gas  turbine  can  be 
operated  by  either  of  two  methods: 

(1 )  By  combustion  of  the  fuel  in  a  chamber  at  constant  pressure. 

(2)  By  an  explosion  method. 

Combus'tion  at  constant  pressure  seems  to  be  the  more  practi- 
cable method  and  is  the  one  generally  adopted.*  In  the  opera  - 

*  Theoretically  the  same  efficiency  should  be  secured  with  either  of  these  two 
systems  of  combustion.  Combustion  at  constant  pressure  is  an  adaptation  of  the 


GAS  TURBINES  349 

tion  of  this  method  gas  and  air  are  compressed  in  separate 
chambers  or  compressor  tanks  to  a  suitable  pressure,  usually 
about  100  pounds  per  square  inch  absolute.  The  gas  and  air 
are  admitted  through  separate  valves  to  the  combustion  chamber, 
where  the  gas  is  ignited  and  burned  at  constant  pressure.  Just 
as  in  a  reciprocating  gas  engine,  the  air  is  provided  to  furnish  the 
oxygen  to  support  combustion.  After  combustion  the  burned 
gases  escape  through  a  suitable  nozzle  to  impinge  on  the  blades 
of  the  turbine  wheel.  On  account  of  the  extremely  high  temper- 
atures resulting  from  the  combustion  (about  2500  degrees  F.), 
it  is  impracticable  to  design  a  gas  turbine  with  more  than  one 
pressure  stage  and  therefore  only  "nozzle  types"  can  be  used. 

Comparison  of  Losses  in  a  Gas  Turbine  and  in  a  Gas  Engine. 
It  is  reasonable  to  assume  that  the  radiation  and  cooling  water 
losses  will  be  about  the  same  for  a  gas  turbine  as  for  a  recipro- 
cating gas  engine;  and  from  a  practical  viewpoint  the  work 
required  for  the  compression  of  gas  and  air  is  about  the  same 
for  combustion  at  constant  pressure  as  for  explosion.  After 
eliminating,  therefore,  the  radiation,  cooling  water,  and  com- 
pression losses  the  same  energy  remains  for  utilization  in  each  of 
these  two  prime  movers.  In  gas  engines  from  20  to  25  per  cent, 
of  this  energy  is  lost  in  the  suction  and  exhaust  resistances,  engine 
friction,  and  the  heat  loss  in  the  exhaust.  Corresponding  to  these 
losses  in  the  gas  engine,  there  are  in  the  gas  turbine  losses  due 
to  nozzle,  blade,  and  disk  friction,  the  heat  in  the  exhaust,  and 
bearing  friction.  The  sum  of  these  latter  losses  in  a  steam  tur- 
bine would  be  about  40  per  cent.,  and  they  will  probably  be  not 
much  different,  in  the  total,  in  a  gas  turbine.  It  is  argued  in 
favor  of  the  gas  turbine  that  it  is  not  impossible  "to  isolate  the 

well-known  Brayton  cycle.  It  has  been  shown  that  exactly  the  same  thermal 
efficiency  can  be  secured  by  such  combustion  as  in  the  ordinary  explosion  process 
if  it  is  assumed  that  the  specific  heat  of  the  gases  is  practically  constant  and  that  the 
final  pressure  after  compression  in  the  explosion  motor  is  the  same  as  the  constant 
pressure  of  combustion  in  the  Brayton  cycle.  It  followsvthen  that  the  ideal  gas  tur- 
bine will  theoretically  operate  with  the  same  fuel  consumption  per  unit  of  power 
as  the  ideal  four-cycle  gas  engine.  (Cf.  Lorenz,  ZeiL  Venin  deutscher  Ingenieure, 
1900,  page  252.) 


350  THE  STEAM  TURBINE 

combustion  chamber  internally"  so  that  no  cooling  water  will 
be  needed. 

The  greatest  practical  difficulty  in  the  way  of  the  successful 
operation  of  gas  turbines  results  from  the  high  temperature  of  the 
gases  at  end  of  expansion.  Nozzles  can  be  cooled  by  water- 
jacketing,  but  the  wheel  blades  are  liable  to  rapid  deterioration. 
The  necessity  for  lowering  the  temperature  of  combustion  is  now 
generally  recognized,  and  the  hot  water  from  the  water  jackets  is 
sprayed  into  the  compressed  air  supplied  for  combustion.  By 
this  means  the  temperature  in  the  combustion  chamber  can  be 
greatly  reduced  but  at  a  considerable  loss,  however,  in  efficiency. 
It  is  making,  in  other  words,  the  efficiency  of  the  gas  turbine 
approach  the  lower  thermal  efficiency  of  the  steam  turbine. 

It  seems  probable  that  the  most  promising  field  for  the  gas 
turbine  will  be  found  to  be  in  the  utilization  of  bituminous  coals 
forming  tar  and  asphaltum  when  used  for  making  gas.  Such 
coals  cannot  be  used  in  the  manufacture  of  gas  for  reciprocating 
gas  engines,  as  the  accumulation  of  tarry  matter  in  the  cylinders 
is  particularly  objectionable.  In  a  gas  turbine,  however,  the  gas 
is  burned  under  pressure,  in  an  enclosed  chamber  where  accumu- 
lations of  foreign  matter  cause  no  serious  difficulties. 

It  cannot  be  expected  that  gas  turbines  can  be  commercially 
successful  for  general  power  purposes  if  a  reciprocating  compressor 
must"  be  used  in  connection  with  them,  because  a  gas  turbine  with 
a  compressor  of  this  type  is  quite  as  complicated  as  the  recip- 
rocating gas  engine.  Compressors  of  the  rotary  type,  on  the  other 
hand,  have  usually  a  very  low  efficiency;  probably  in  most  cases 
not  more  than  50  per  cent.  Decided  progress  is  being  made  in 
the  successful  designing  of  compressors  of  the  turbine  type  which 
will  give  from  60  to  70  per  cent,  efficiency.  It  is  not  difficult  to 
understand  how  the  net  useful  work  of  a  gas  turbine  may  be  nil 
under  conditions  that  are  not  unusually  poor.  For  if  the  effi- 
ciency of  the  turbine  is  60  per  cent,  and  the  theoretical  work  of 
compression  is  40  per  cent,  of  the  output  (which  is  not  an  absurd 
estimate),  then  with  a  compressor  efficiency  of  only  40  per  cent, 
the  theoretical  power  absorbed  by  the  compressor  is  60  X  .40,  or 


GAS  TURBINES 


351 


24  per  cent,  of  the  output,  or  the  actual  power  delivered  to  the 
compressor  is  24  •+•  .40,  or  60  per  cent,  of  the  output.  And  the 
compressor  takes  all  the  power  the  turbine  can  supply.  It  is 
obvious  then  that  compressors  with  the  usual  low  efficiencies  of 
the  rotary  types  are  not  worth  considering. 

BRAYTON    CYCLE    CALCULATIONS    FOR    GAS    TURBINES. 


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Case  I: 

90 

1,000 

0 

4,665 

2,435 

.22 

115,300 

43 

Adiabatic    com- 
pression. Per- 
fect   machine 
with  no  losses. 

90 
195 
195 

495 
495 

250 

1,000 

250 

1,000 

250 

0 
0 
0 
0 
0 

i,5°5 
4,860 
1,710 
5,190 
2,040 

652 
2,008 
546 
1,562 

434 

.87 
.27 
1.07 

•34 
1-38 

71,680 
131,200 
83,740 
147,600 
98,300 

43 
54 
54 
64 
64 

Case  II: 

Isothermal  com- 

90 

1,000 

0 

9,239 

5,039 

.08 

159,100 

93 

pression  with 

90 

250 

0 

1,965 

•39 

79,57o 

72 

regenerator. 

195 

1,000 

0 

7,376 

3,176 

.11 

159,100 

91 

Perfect      ma- 

195 

250 

0 

1,498 

448 

.61 

79,570 

62 

chine  with  no 

495 

1,000 

0 

6,087 

1,887 

.  15 

159,100 

87 

losses. 

495 

250 

0 

1,176 

126 

1.03 

79,57o 

49 

Case  III: 

Isothermal  com- 

56 

333 

0 

2,147 

1,200 

•74 

75,4oo 

27 

pression  with 

82 

392 

0 

2,398 

1,200 

.68 

84,800 

3° 

regenerator. 

61 

445 

0 

2,867 

1,619 

•50 

86,485 

Actual       ma- 
chine with  as- 
sumed losses. 

98 

33 
61 

545 
445 
569 

0 
0 

o 

3,233 

1,619 
2,!39 

•47 
.40 

•36 

99,96o 
80,155 
96,775 

35 

28 

33 

Air  excess. 

Case  IV: 

Isothermal  com- 

pression   with 

79 

625 

•37 

1,275 

600 

•67 

75,000 

13 

regenerator. 

47 

638 

•36 

i,  680 

1,000 

•43 

75,000 

15 

Assumed  losses  . 

Cooling  water. 

352  THE   STEAM  TURBINE 

THERMODYNAMIC   THEORY   OF    THE   GAS   TURBINE. 

The  elementary  thermodynamics  of  the  gas  turbine  involve 
apparently  no  new  investigations.  The  problems  are  princi- 
pally mechanical  and  metallurgical.  In  the  above  table  the 
efficiencies  of  various  gas  turbine  cycles  are  given  as  calculated 
by  Sanford  A.  Moss.  In  all  cases  it  is  assumed  that  the  heat 
of  combustion  is  developed  at  constant  pressure  and  that  the 
exhaust  gases  are  discharged  at  constant  pressure.  Cases  I  and 
II  refer  to  theoretically  perfect  engines,  and  cases  III  and  IV  to 
engines  with  probably  normal  losses.  It  is  assumed  for  these 
latter  cases  that  the  turbine  efficiency  is  70  per  cent,  and  of  the 
compressor  is  83  per  cent.*  The  efficiency  of  the  regenerator 
used  for  the  cases  of  isothermal  compression  is  taken  to  be  60 
per  cent.  These  figures  are  certainly  above  the  upper  limits  of 
possible  results  in  practice. 

Thermodynamic  efficiencies  of  gas  turbines  operating  with 
combustion  at  constant  pressure  will  now  be  discussed.  The 
equations  given  are,  in  most  cases,  those  relating  to  a  perfect 
gas. 

The  total  heat  H  of  a  gas  at  constant  pressure  may  be  expressed 
by  the  following  equation: 

H  =  cvT  +  RT  =  (ctf  +  R)  T  =  cpT  +  constant, 

where  T  is  the  absolute  temperature,  cv  and  cp  are  respectively 
the  mean  specific  heats  of  the  gas  at  constant  volume  and  con- 
stant pressure  between  zero  temperature  and  T,  and  R  is  a  con- 
stant varying  for  its  value  with  the  kind  of  gas. 

Fig.  225  represents  the  cycle  of  operations  when  a  pound  of  a 
mixture  of  gas  and  air  is  compressed  and  later  expanded  in  doing 
work.  Adiabatic  compression  is  assumed.  One  pound  of  the 
mixture  is  taken  into  the  compressor  cylinder  at  the  temperature 
Tt  and  volume  v0  and  is  compressed  as  represented  by  the  adia- 
batic  O  3  to  the  temperature  T3  and  volume  v3.  In  the  passage 
to  the  combustion  chamber  it  will  be  assumed  for  simplicity  in 

*  Efficiencies  of  60  per  cent,  for  the  turbine  and  not  more  than  70  per  cent,  for 
the  compressor  would  probably  be  more  reasonable. 


GAS  TURBINES 


353 


the  calculations  that  the  temperature  drops  to  the  initial  temper- 
ture  T0.  If  the  total  heat  contents  at  the  points  o,  3,  3',  4,  and 
5  are  represented  by  the  corresponding  symbols  H0,  H3,  H3',  H4, 


Volume 


FIG.  225.     Diagram  of  the  Theoretical  Action  of  a  Gas  Turbine  and 
Air  Compressor. 


and  H5,  the  indicated  work  of  compression  is  the  area  0123, 
which  will  be  represented  in  heat  units  by 

W—  II  IT 

c    —  ±13   —    ±10. 

It  is  assumed,  however,  that  immediately  after  compression  the 
temperature  falls  to  T0,  so  that  the  volume  is  reduced  from  v3  to 
v/.  Now  because  the  point  3'  is  on  the  isothermal  0  3',  it  is 
obvious  that  H0  =  H3'.  During  combustion  a  quantity  of  heat 
Qt  is  added  to  the  mixture,  increasing  the  temperature  to  T4  and 
the  volume  to  v4;  or,  in  other  words, 

Q,-H4-H,'. 

When  the  gaseous  mixture  is  expanded  the  work  Wl  is  per- 
formed, which  may  be  calculated  after  determining  H4  in  the 
preceding  equation;  then 

Wt=H4-H5. 


354  THE  STEAM  TURBINE 

The  heat  lost  by  the  exhaust  gases  in  cooling  from  T5  to  T0  is 
H5  —  H0;  and  since  this  quantity  is  called  Q2,  we  can  write 

W,  -  (H4  -  H,')  -  (H,  -  H.)  =  Q,  -  Q2, 
and  placing  H0  for  H3'  it  is  apparent  that 

W,  =H4-H.=  0,-Q,. 

The  theoretical  discharge  velocity  V  in  feet  per  second  at 
the  mouth  of  the  expansion  nozzle  is  calculated  from  the  usual 
equation, 

V2 


2g   X    778 


=  Ql-Q,=  H4-H5  =c,(T4-T5). 


Using  the  same  symbols  as  before  for  the  indicated  work  of 
compression,  the  theoretical  effective  power  of  the  turbine  is 

We  =  W,  -  Wc  =  Q1  -  Q2  -  Wc. 

If  the  mechanical  efficiency  of  compression  is  x  and  the  efficiency 
of  the  gas  turbine,  y,  is  determined  in  the  same  way  as  for  a  steam 
turbine,  by  constructing  velocity  triangles  and  calculating  the 
nozzle,  blade,  and  wheel  friction  losses  for  a  single  stage  turbine, 
then  the  theoretical  net  power  of  the  turbine  is 

W.'  =  (Q,  -  Q2)  y  -  ^- 

Since  the  heat  consumed  per  pound  of  the  mixture  is  Q1?  the 
total  efficiency  z  of  the  gas  turbine  apparatus  is 


((Q,  -  Q2)y  -  Y)-" 


Efficiency  of  Gas  Turbine  with  Water  Injection.  If  m  pounds 
of  water  are  injected  into  the  combustion  chamber  just  before 
the  expansion  begins,  an  equal  weight  of  steam  is  found,  which 
it  will  be  assumed  is  superheated  to  the  temperature  T4',  which 
will  now  be  also  at  the  temperature  of  the  mixture,  lower,  of 
course,  thai)  T4. 


GAS  TURBINES  355 

The  temperature  of  the  mixture  of  burned  gases  and  steam 
T4'  is  calculated  by  solving  the  following  equation: 

c,  (T,  -  T/)=  m  }q/  -  q,  +  r/  +  c/  (t/  -  tS4)j, 

where  q/  is  the  heat  of  the  liquid,  r/  is  the  heat  of  vaporization, 
and  tS4  is  the  temperature  of  saturated  steam,  —  all  at  the 
corresponding  pressure  P/.*  The  other  new  symbols  are  q;, 
which  is  the  heat  of  the  liquid  at  the  injection  temperature,  and 
c/,  which  is  the  specific  heat  of  superheated  steam.  In  this 
equation  t/  and  ts4  are  ordinary  (not  absolute)  temperatures. 

The  temperature  T5'  is  calculated  for  assumed  adiabatic  expan- 
sion by  using  the  exponent  k'  calculated  from  the  equation 
below  : 


cvi  4- 

in  which  the  subscript  i  refers  to  the  specific  heats  of  the  mixture 
and  the  subscript  2  to  specific  heats  of  the  steam.  The  temper- 
ature T5'  is  used  to  determine  the  value  of  Q2?  which  is  the  quan- 
tity of  heat  abstracted  from  the  mixture  to  cool  it  from  the 
condition  at  5  to  the  condition  at  o.  It  is  calculated  from  the 
following  equation: 

Q2  =  m  icp'  (t/  -tS5)  +  r5'  +  q/  -  q0l  +  cp(t5'  -t0). 

In  this  equation  t5'  is  the  "  ordinary"  temperature  corresponding 
to  the  absolute  temperature  T/,  tS5  is  the  temperature  of  satu- 
rated steam,  r5'  is  the  heat  of  vaporization,  and  q/  is  the  heat  of 
the  liquid,  —  all  at  the  pressure  p5';  q0  is  at  the  temperature  t0.  It 
will  be  assumed  also  that  t0  is  less  than  tS5  and  that  the  latter  is 

*  The  "partial"  pressure  of  the  steam  at  the  temperature  T4'  can  be  calculated 
approximately  by  the  formula 

p  i  _    .  47  »• 

•*      A  ' 


29.3  +  47  m 

if  we  assume  the  constants  for  the  exhaust  gases  are  the  same  as  for  air.     In  the 
same  way  the  "partial"  pressure  of  the  steam  after  expansion  is  calculated  thus: 


r.'-p, 


29.3  4-  47  m 


356  THE  STEAM  TURBINE 

less  than  t5',  as  is  generally  the  case,  neglecting  the  small  quan- 
tity of  heat  in  the  water  vapor  remaining  in  the  mixture  at  the 
temperature  t0  after  the  burned  gases  and  steam  have  been  dis- 
charged from  the  nozzle. 

We   can   write   the   following   equations,  applying  the   same 
methods  as  for  the  case  without  the  use  of  injection  water: 

Y.'d+m) 

2g  X   778 


Total  efficiency,  z,  increases  with  the  pressure  of  compression  to 
a  certain  limiting  value  and  then  decreases.  But  for  the  practi- 
cable values  of  compressor  and  turbine  efficiencies  (x  and  y)  the 
values  of  the  theoretical  total  efficiency  are  not  particularly  good. 

The  equations  given  here  for  velocity  and  efficiency  can  be 
used  to  investigate  the  best  operating  conditions  by  varying  the 
pressure  of  compressions  and  the  quantity  of  injection  water. 

Another  method  for  reducing  the  temperature  of  the  gases  is 
to  use  a  large  excess  of  air  above  the  quantity  needed  to  support 
combustion.  The  most  economical  method  is  probably  that  of 
partly  vaporizing  the  cooling  water  in  the  water  jackets.  The 
advantage  of  using  water  is  that  likewise  it  does  not  need  to  be 
compressed,  and  therefore  it  can  be  injected  into  the  combustion 
chamber  without  the  expenditure  of  much  energy. 

Many  of  the  troubles  in  the  combustion  chamber  of  a  gas 
turbine  are  difficult  to  explain.  One  of  the  most  serious  diffi- 
culties is  the  occasional  missfire  of  the  incoming  charge  which  is 
soon  followed  by  a  violent  explosion.  It  is  also  difficult  to  secure 
smokeless  combustion.  Improvements  are  being  made,  how- 
<ever,  with  the  object  of  maintaining  higher  temperatures  in  the 
•combustion  chamber,  and  the  results  are  encouraging.  For  this 
reason  it  is  important  that  cooling  water  should  be  injected  into 


GAS  TURBINES  357 

the  gases  after  they  have  left  the  chamber.  The  use  of  carbo- 
rundum for  lining  the  combustion  chamber  and  for  the  nozzles  is 
apparently  an  important  step  forward.  This  material,  which  is 
a  product  of  the  electric  furnace,  is  therefore  manufactured  at  a 
much  higher  temperature  than  is  ever  attained  in  a  gas  com- 
bustion chamber. 


CHAPTER   XVII. 
ELECTRIC  GENERATORS  FOR  STEAM  TURBINES. 

IN  the  early  years  of  the  development  of  steam  turbines  it  was 
the  primary  aim  of  the  turbine  engineer  to  reduce  the  speed  of 
the  turbine  to  operate  satisfactorily  when  direct  connected  to 
electric  generators.  To  accomplish  this  purpose  De  Laval 
introduced  his  famous  helical  gearing  and  Curtis  applied  the 
principle  of  velocity  stages.  To-day,  however,  the  trend  of 
developments  is  in  the  other  direction.  The  electrical  engineer 
is  being  urged  to  use  his  best  skill  to  design  generators  to  operate 
satisfactorily  at  higher  and  higher  speeds,  because  in  this  way 
the  efficiency  of  the  turbine  can  best  be  increased.  Very  great 
strides  have  been  made  in  the  perfection  of  alternators  for  steam 
turbines;  but  in  the  design  of  direct-current  generators  to  operate 
at  high  speeds  much  is  still  to  be  desired.  In  fact,  for  high 
speeds,  commutation  is  indeed  a  very  difficult  problem.  The 
potential  difference  between  adjacent  commutator  bars  of  engine- 
driven  direct-current  generators  is  usually  about  10  volts;  while 
for  turbine-generators  the  limit  is  about  30  to  40  volts  per  bar. 

DIRECT-CURRENT   GENERATORS. 

Sparking  Limit.  In  slow-speed  direct-current  generators  the 
output  is  limited  either  by  the  sparking  or  by  the  heating.  On 
account  of  the  necessarily  small  dimensions  of  the  armature  and 
the  extremely  high  periodicity,  with  the  resulting  large  iron  losses, 
it  becomes  necessary,  on  the  other  hand,  in  a  high-speed  generator 
to  employ  artificial  cooling  devices.  With  the  use  of  forced 
ventilation  heating  is  no  longer  a  factor  limiting  the  output,  and 
the  difficulty  lies  then  principally  in  the  sparking.  The  quality 
of  the  commutation  in  any  electric  generator  depends  largely  on 

358 


ELECTRIC  GENERATORS  FOR  STEAM  TURBINES     359 

the  number  of  ampere-turns  which  can  be  placed  on  the  surface 
of  the  armature;  but  there  are  also  a  number  of  other  electric 
and  magnetic  conditions  to  be  considered,  particularly  the  effec- 
tiveness of  the  commutating  poles,*  located  between  the  main 
poles.  Another  important  factor  is  the  mechanical  condition 
of  the  armature,  commutator,  and  brushes,  which  determines  in 
a  large  measure  the  sparking  limit. 

On  the  basis  that  the  maximum  permissible  ampere-turns  per 
centimeter  of  the  circumference  of  the  armature  determine  the 
maximum  output  the  following  table  has  been  calculated.  At 
an  assumed  permissible  peripheral  velocity  of  75  meters  f  per 
second  J  (about  245  feet  per  second),  this  table  shows  the  maxi- 


Revolutions 

Diameter  of 

Output 

per 

Armature  in 

in 

Minute. 

Centimeters  . 

Kilowatts. 

4780 

3° 

148 

2870 

5° 

347 

1800 

80 

670 

1435 

100 

890 

895 

160 

•  1585 

720 

200 

2080 

*  When  artificial  commutation  is  secured  by  auxiliary  poles  placed  between  the 
main  poles  of  a  generator,  short-circuit  currents  and  sparking  can  occur  only  when 
the  electromotive  force  induced  by  the  commutating  field  is  different  from  the 
reactance  voltage 

f  The  C.  G.  S.  (metric)  system  of  units  is  applied  in  this  chapter  because  it  is 
the  one  commonly  used  by  designers  of  electrical  machinery  in  America  and 
in  England. 

J  It  is  usually  found  that  the  end  shells  or  shields  protecting  the  connections  in 
a  revolving  armature  have  stresses  most  nearly  approaching  the  allowable  limits. 
Stresses  in  these  end  shells  are  calculated  as  in  a  ring  or  band,  by  equation  (29), 
which  becomes  approximately  in  C.  G.  S.  units, 

_ 

where  V  is  the  peripheral  velocity  of  the  ring  in  meters  per  second,  2  is  the  weight 
of  a  cubic  centimeter  of  the  material,  and  Sa  is  the  allowable  unit  stress  in  kilograms 
per  square  centimeter.  Since  the  allowable  permissible  stress  of  bronze  castings  is 
about  260  kilograms  per  square  centimeter,  the  maximum  allowable  velocity  is  only 


360  THE  STEAM  TURBINE 

mum  outputs  and  speeds  which  at  the  present  time  are  obtain- 
able in  the  very  best  designs  of  direct-current  generators.* 

This  table  shows  that  the  armature  of  direct-current  generators 
cannot  be  constructed  to  give  the  required  output  at  the  usual 
speeds  adopted  in  America  for  Parsons  turbines,  and  that  with 
Curtis  turbines,  which  operate  at  slower  speeds,  the  limit  is 
reached  at  1500  kilowatts  full  load  capacity. 

There  are  two  ways  of  overcoming  the  limitations  of  direct- 
current  generators  for  turbine  service.  One  way  is  to  design  the 
turbines  for  lower  speeds,  which  entails,  however,  increased  cost 
and  a  sacrifice  of  economy.  The  other  way  is  to  adopt  the  tan- 
dem arrangement  of  connecting  two  generators  to  one  turbine 
as  in  the  usual  De  Laval  designs  for  the  larger  sizes. 

There  is  a  constant  demand  for  direct-current  generators  of 
larger  capacities  than  are  now  employed,  and  the  problem  of 
increasing  the  capacity  of  the  generator  is  becoming  very  impor- 
tant. The  successful  production  of  such  machines  suitable  for 
much  higher  speeds  than  are  now  attainable  would  be  an  improve- 
ment effective  in  two  ways:  (i)  by  lowering  the  steam  consump- 
tion and  (2)  by  reducing  the  first  cost;  and  as  a  result  the  field 
of  the  high-speed  reciprocating  engine  would  be  still  more 
restricted. 

Flash-over  Limit.  In  ordinary  slow-speed  direct-current 
generators  the  only  electrical  limit  to  the  capacity  is  sparking. 
In  high-speed  machines,  however,  a  new  difficulty  known  as  the 
flash-over  limit  is  met.  Its  effects  are  often  as  serious  and  as 
difficult  to  remedy  as  any  of  the  commutation  troubles.  A  great 
many  designs  of  high-speed  direct-current  generators  with 
satisfactory  commutating  qualities  have  been  failures  because 
of  their  tendency  to  arc  around  the  whole  commutator.  This 
trouble  must  be  attributed  primarily  to  the  very  high  potential 

55  meters  per  second;  and  considering  the  additional  load  due  to  the  end  connec- 
tions the  permissible  velocity  becomes  only  about  50  meters  per  second.  If,  how- 
ever, phosphor-bronze  or  manganese-bronze  castings  with  an  allowable  stress  of 
600  kilograms  per  square  centimeter  are  used,  a  peripheral  velocity  of  75  meters 
per  second  is  not  excessive. 

*  R.  Pohl,  Proc.  of  Inst.  of  Elec.  Engrs.,  1907. 


ELECTRIC  GENERATORS  FOR    STEAM  TURBINES      361 

difference  between  adjacent  commutator  bars  —  usually  about 
three  times  the  permissible  value  in  slow-speed  generators. 
Usually  this  difficulty  can  be  remedied  by  increasing  the  insula- 
tion of  the  shrinkages  and  of  the  brush-gear.  It  is  the  flash- 
'  over  limit,  therefore,  which  determines  the  allowable  voltage 
per  commutator  bar  and  restricts  the  number  of  "lines"  or  flux 
allowed  to  enter  or  leave  an  armature  of  a  given  diameter. 

The  most  obvious  line  of  improvement  in  turbine-driven 
direct-current  generators  is  in  increasing  the  peripheral  speed  of 
the  armature.  Steel  alloys  of  very  low  magnetic  conductivity 
and  high  tensile  strength  used  in  the  place  of  phosphor-bronze 
for  the  end  shields  of  the  armature  will  permit  the  adoption  of 
considerably  higher  peripheral  speeds  than  are  now  allowable. 
If  we  compare  two  machines  of  equal  output  and  speed  but  with 
armatures  of  different  diameters  in  the  ratio  of  i  to  2,  the  arma- 
ture with  the  larger  diameter  will  be  only  one-fourth  as  long  as 
the  other;  while  with  the  same  voltage  for  both  the  number  of 
conductors  and  segments  will  be  doubled.* 

ALTERNATING-CURRENT  GENERATORS. 

The  design  of  alternators  with  revolving  fields  to  operate  at 
high  speeds  is  not  nearly  so  difficult  as  for  commutating  machines. 
Speed  limits  are  usually  determined  by  the  strength  of  suitable 
materials  for  their  construction.  When  it  became  the  general 
practice  to  enclose  high-speed  generators  in  a  sheet-metal  casing 
and  to  adopt  forced  or  artificial  ventilation  produced  by  small 
fans  circulating  air  through  the  generator  windings,  it  was 
possible  to  regulate  the  heating  limit  so  that  heavier  overloads 
could  be  carried  and  for  longer  periods  of  time  than  was  possible 
before.  By  this  method  the  excessive  noise  of  the  early  turbine- 
generators  was,  at  the  same  time,  eliminated. 

For  the  windings  of  the  revolving  fields  of  high-speed  alterna- 
tors, flat  strap  copper  is  used  by  most  manufacturers.  To  make 
the  field  spools  as  small  and  compact  as  possible  this  strap  copper 

*  The  voltage  per  segment  is  approximately  inversely  proportional  to  the 
peripheral  velocity  of  the  armature. 


362 


THE  STEAM  TURBINE 


is  sometimes  coated  with  a  thin  layer  of  enamel  for  insulation, 
instead  of  the  usual  cotton  covering.  The  necessity  of  making 
the  exterior  surfaces  of  revolving  fields  as  smooth  as  possible  is 
generally  appreciated  by  designers. 


GENERATOR  EFFICIENCIES. 


Average  efficiencies  of  the  best  designs  of  alternating-current 
generators  intended  for  operation  with  steam  turbines  are  given 
in  the  following  table : 


Rated  Full  Load  Capacity, 
Kilowatts. 

Efficiency  of 
Alternator, 
Per  Cent. 

50  to  150 

9°-93 

200  tO  400 

93-94 

500  to  900 

95 

1000  to  2500 

96 

3000  to  5000 

97 

6000  to  10,000 

98 

The  efficiency  of  direct-current  high-speed  generators  is  about 
one  per  cent,  less  than  that  of  alternators  of  the  same  capacity. 


APPENDIX. 
EXERCISES  ON  STEAM  TURBINES. 

Exercise  i.  What  is  the  velocity  of  steam  discharging  at  the  rate  of  200 
cubic  feet  per  second  through  a  nozzle  having  a  cross-sectional  area  of  0.2 
square  foot?  Ans.  1000  feet  per  second. 

Exercise  2.  If  the  steam  discharging  from  the  orifice  mentioned  in  the 
preceding  exercise  weighs  .0322  pound  per  cubic  foot,  how  much  energy 
in  foot-pounds  per  second  can  this  jet  develop  ?  How  much  horsepower  ? 
Ans.  100,000  foot-pounds  per  second.  1  8  1.8  horsepower. 

Suggestion:  From  elementary  mechanics  we  have  the  information  that 
the  kinetic  energy  K  (sometimes  called  capacity  to  do  work)  of  any  moving 
fluid,  such  as  steam,  gas,  or  water,  is 


where  W  is  the  weight  of  the  fluid  discharging  per  second,  V  is  the  velocity 
of  flow  in  feet  per  second,  and  g  is  the  acceleration  of  gravity  or  32.2  feet 
per  second. 

By  definition  (in  English  units)  one  horsepower  is  equivalent  to  550 
foot-pounds  per  second. 

Exercise  3.  If  the  vessel  shown  in  Fig.  33  discharges  40  pounds  of  water 
per  second  at  a  velocity  of  161  feet  per  second,  what  is  the  force  (impulse) 
pushing  the  wooden  block  away  from  the  vessel  ?  Ans.  200  pounds. 

Also  what  is  the  force  (reaction)  pushing  the  vessel  itself  toward  the 
left?  Ans.  200  Ibs. 

Exercise  4.  If  water  is  discharged  against  flat  blades  of  a  water  wheel 
made  up  of  vanes  similar  to  the  block  shown  in  Fig.  33  (page  58)  at  the 
rate  of  3.22  pounds  per  second  at  a  velocity  of  2000  feet  per  second  and  is 
spattered  from  the  wooden  blocks  with  a  "residual"  velocity  (leaving  the 
vanes)  of  300  feet  per  second,  what  horsepower  is  this  water  wheel  capable 
of  developing?  Ans.  195,500  foot-pounds  per  second  or  3555  horsepower. 

Exercise  5.  Steam  of  the  same  density  as  in  the  exercise  on  page  28 
discharges  at  the  rate  of  1739  pounds  per  hour  and  produces  a  reaction 
against  the  plate  into  which  the  same  nozzle  is  inserted  of  45  pounds.  What 
is  the  velocity  of  discharge?  Ans.  3000  feet  per  second. 

363 


364  APPENDIX 

Exercise  6.  The  area  of  a  nozzle  at  its  smallest  section  is  .72  square  inch 
and  discharges  steam  at  the  rate  of  .2  pound  per  second,  of  which  the  specific 
volume  is  2.0  cubic  feet  per  pound. 

(a)  What  is  the  velocity  of  flow?  Ans.  80  feet  per  second. 

(6)  What  is  the  magnitude  of  the  force  developed  by  the  reaction  of 
this  jet?  Ans.  \  pound  (nearly). 

(c)  What  is  the  maximum  value  of  the  impulse  produced  by  this  jet  if 
friction  and  eddy  losses  reduce  the  velocity  effective  for  giving  the  impulse 
by  25  per  cent.?  Ans.  .37  pound. 

(d)  If  only  part  of  the  velocity  available  in  (c)  is  absorbed  in  driving  a 
steam  turbine,  so  that  the  steam  leaves  the  blades  with  a  "residual"  veloc- 
ity of  10  feet  per  second,  how  many  foot-pounds  of  work  per  minute  are 
developed  by  the  turbine  ?  Ans.  652. 

(e)  What  is  the  horsepower  equivalent  of  this  number  of  foot-pounds 
per  minute?  Ans.  .0198  horsepower. 

(/)  If  this  turbine  drives  a  small  electric  generator  having  an  efficiency 
of  80  per  cent.,  what  power  in  kilowatts  will  this  generator  develop  ? 

Ans.  .0133  kilowatt. 

Suggestion:  A  kilowatt  is  a  thousand  watts,  and  746  watts  are  equiva- 
lent to  a  horsepower. 

(g)  How  much  horsepower  would  be  developed  by  this  turbine  if  all  the 
velocity  as  calculated  in  (a)  is  transformed  into  work  ? 

(ti)  What  is  the  efficiency  of  the  turbine? 

Suggestion:  Compare  (e}  and  (g).  If  the  velocity  as  calculated  in  (a) 
represents  the  total  velocity  equivalent  of  the  available  energy  due  to 
adiabatic  expansion  (constant  entropy)  then  the  answer  to  section  (ti)  is 
called  the  Rankine  efficiency  of  the  turbine. 

Exercise  7.  Calculate  the  horsepower  developed  by  a  steam  turbine 
having  two  rows  of  moving  blades.  Upon  the  first  row  steam  is  directed 
at  a  velocity  of  3000  feet  per  second  and  at  the  rate  of  1.771  pounds  per 
second.  The  steam  is  discharged  from  this  row  at  a  velocity  of  1000  feet 
per  second  and  is  then  directed  upon  a  second  row  of  blades  from  which  it 
is  discharged  at  a  velocity  of  200  feet  per  second. 

(a)  Neglecting  frictional  and  other  losses,  how  much  horsepower  will 
this  turbine  develop  ?  Ans.  448  horsepower. 

(b)  How  much  power  would  be  developed  if  there  is  a  loss  of  velocity  of 
10  per  cent,  in  each  row  of  blades? 

Suggestion:  Actual  velocity  effective  in  first  row  of  blades  is  2700  feet 
per  second  which  is  discharged  at  1000  feet  per  second.  Work  developed 
in  this  row  is  1.771  (27oo2  —  iooo2)  -5-  (64.4  X  550)  in  horsepower.  Simi- 
larly work  done  in  the  second  row  is  1.771  (poo2  —  2oo2)  •*-  (64.4  X  550) 
in  horsepower. 


APPENDIX  365 

(c)  What  is  the  efficiency  of  the  complete  turbine  when  the  losses  stated 
are  considered? 

Suggestion:  Efficiency  is  total  horsepower  calculated  in  (b)  divided  by 
that  found  by  considering  only  initial  velocity  (3000  feet  per  second)  as 
in  (g)  of  Exercise  6. 

Exercise  8.     (Use  of  entropy-total  heat  chart.) 

Steam  at  an  initial  condition  of  165  pounds  per  square  inch  absolute 
and  100  degrees  Fahrenheit  superheat  is  expanded  in  a  nozzle  adiabatically 
to  20  pounds  per  square  inch  absolute  pressure. 

(a)  How  much  energy  (in  B.T.U.)  is  converted  into  velocity? 

(b)  If  no  losses  are  considered,  what  is  the  velocity  in  feet  per  second  of 
the  discharging  jet  ? 

(c)  If  there  are  losses  equivalent  to  4  per  cent,  of  the  energy  available, 
what  is  the  actual  velocity  of  the  jet  ? 

(d)  If  the  losses  are  equivalent  to  2  per  cent,  of  the  theoretical  velocity, 
what  is  the  actual  velocity  of  the  jet  ? 

Exercise  9.  Steam  at  the  same  initial  and  final  conditions  as  in  Exer- 
cise 8  is  reheated  by  friction  in  the  nozzles  and  blades  so  that  the  entropy 
at  the  final  condition  is  1.7.  How  much  energy  is  available  for  doing  work? 

Suggestion:  Reading  from  the  entropy-total  heat  chart,  the  total  heat 
contents  of  a  pound  of  steam  at  the  initial  condition  is  1242  B.T.U.  and  at 
the  final  condition  after  reheating  is  1135  B.T.U.  (expansion  is  not  adia- 
batic).  Therefore,  heat  units  available  for  work  =1252  —  113 5  or  117  B.T.U. 

Exercise  10.  A  certain  steam  turbine  having  several  stages  takes  steam 
initially  at  165  pounds  per  square  inch  absolute  and  100  degrees  F. 
superheat  and  expands  it  to  20  pounds  per  square  inch  absolute  in  the 
first  stage.  Friction  and  the  transformation  of  residual  velocity  into 
potential  (heat)  energy  returns  30  per  cent,  of  the  available  energy  in  an 
adiabatic  expansion  back  to  the  steam  by  "reheating"  at  the  final  pressure. 
If  now  in  the  nozzles  of  a  succeeding  stage  of  the  turbine  the  steam  is 
expanded  to  lo'pounds  per  square  inch  absolute  and  reheated  again  by  the 
same  percentage  at  the  latter  pressure,  how  much  energy  (B.T.U.)  is  avail- 
able in  each  stage  for  performing  work  ?  What  is  the  quality  of  the  steam 
after  each  reheating? 

Suggestion:  In  the  normal  adiabatic  expansion  from  165  pounds  per 
square  inch  absolute  and  100  degrees  F.  superheat  to  20  pounds  per  square 
inch  absolute  the  available  energy  is  1252  —  1085  or  167  B.T.U.,  of  which 
30  per  cent,  or  50.1  B.T.U.  go  to  reheat  the  steam.  The  total  heat  con- 
tents of  a  pound  of  steam  after  reheating  becomes  then  1085  +  50.1  or 
1135.1  B.T.U.  For  the  second  stage  the  expansion  is  from  20  pounds  per 
square  inch  absolute  pressure  and  a  total  heat  contents  of  1135.1  B.T.U. 
(quality  about  .975)  to  10  pounds  per  square  inch  absolute,  making  the 


366  APPENDIX 

available  energy  for  adiabatic  expansion  1135.1  —  1087  or  48.1  B.T.U. 
per  pound.  The  reheating  is  30  per  cent,  of  this  or  about  14.3  B.T.U.,  which 
when  added  to  1087  gives  1091.3  B.T.U.  as  the  total  heat  contents  of  the 
steam  when  passing  into  the  nozzles  of  the  next  succeeding  stage.  At  this 
condition  the  quality  of  the  steam  is  about  .949. 

In  actual  designing  the  reheating  in  the  last  turbine  is  not  considered 
available,  as  will  be  observed  in  the  design  worked  out  on  page  87.  The 
reason  for  this  is  that  a  very  large  part  of  the  reheating  in  the  stages  other 
than  the  last  is  due  to  the  changing  of  the  residual  velocity  of  the  steam 
as  it  leaves  the  blades  into  potential  (heat)  energy.  This  has  been  dem- 
onstrated by  actual  experiments  which  show  that  the  steam  enters  the 
nozzle  of  an  impulse  wheel  in  every  stage  with  practically  negligible  velocity. 
In  the  last  stage  the  conditions,  however,  are  different.  The  steam  here 
leaves  the  blade  with  its  residual  velocity  unchecked  and  passes  off  into 
the  large  exhaust  passages  provided  for  its  unimpeded  flow. 

Exercise  u.  A  turbine  blade  like  the  one  shown  in  Fig.  43  moves  with 
a  velocity  of  500  feet  per  second  due  to  a  steam  jet  passing  over  it  which 
has  a  velocity  of  3220  feet  per  second.  If  friction  losses  in  the  blade  are 
not  considered,  and  the  weight  of  steam  flowing  per  second  is  1.0642  pounds, 
a  is  20  degrees  and  /3  =  7  =  45  degrees,  what  is  the  total  impulse  force 
to  which  the  blade  is  subjected  in  the  direction  of  its  motion  (see  page  70)? 

Suggestion:  Since  losses  in  the  blades  are  neglected  Fr3  =  Fr2  and  Fr2  = 
v'Fb2  +  F22  -  2  F6F2coso<  (Law  of  Cosines). 

All  the  terms  in  this  equation  are  known  so  that  Fr2  or  Fr3  can  be  calcu- 
lated. 

Exercise  12.  Taking  the  necessary  data  from  the  preceding  exercise, 
state  the  proper  angle  for  the  backs  of  the  blades  (see  Figs.  49  and  50), 
for  the  steam  to  enter  without  loss  due  to  impact  and  eddying.  (See  page 
68.) 

Exercise  13.  Explain  the  essential  principle  of  operation  of  Hero's 
engine.  Indicate  clearly  in  a  figure  the  direction  of  rotation  of  this  engine 
with  respect  to  the  direction  of  steam  discharge  from  the  nozzles.  What 
is  the  difference  in  principle  between  Hero's  engine  and  Branca's? 

Exercise  14.  Explain  the  actual  difference  between  the  commercial  types 
known  as  impulse  and  reaction  turbines. 

Exercise  15.  Why  have  the  stationary  blades  or  buckets  shown  in 
figures  like  39,  page  63b,  a  curvature  in  the  opposite  direction  to  that  of  all 
the  moving  blades  ? 

Exercise  16.  Why  is  the  rotation  loss  when  stated  in  per  cent,  of  rated 
output  less  for  a  large  size  turbine  than  for  a  relatively  small  one? 

Exercise  17.  Design  a  nozzle,  showing  all  the  important  dimensions, 
for  expanding  steam  from  the  initial  condition  of  165  pounds  per  square 


APPENDIX  367 

inch  absolute,  and  100  degrees  F.  superheat  to  a  final  condition  of  4  pounds 
per  square  inch  absolute.  Assume  that  the  nozzle  loss  is  3  per  cent,  of  the 
velocity  and  that  the  rate  of  flow  is  to  be  ^  pound  per  second.  (See  pages 

93-95-) 

Exercise  18.  Steam  expands  in  the  nozzles  of  a  simple  impulse  turbine 
from  165  pounds  per  square  inch  absolute  to  i  pound  per  square  inch 
absolute  (about  28  inches  vacuum).  Draw  velocity  diagrams,  allowing  for 
no  losses,  and  determine  the  proper  blade  angles  when  ft  equals  y.  The 
nozzle  angle  is  to  be,  as  usual,  20  degrees  and  the  peripheral  speed  of  the 
blades  or  buckets  is  1200  feet  per  second. 

Calculate  the  energy  absorbed  or  given  up  to  the  blades  or  buckets  per 
pound  of  steam  as  well  as  the  steam  consumption  of  the  ideal  turbine 
(theoretical  water  rate)  and  the  steam  consumption  of  this  turbine  as  de- 
termined from  the  energy  absorbed  by  the  blades  or  buckets. 

Sketch  with  a  reasonable  degree  of  accuracy  the  outlines  of  the  blades  or 
buckets. 

Exercise  19.  Recalculate  and  redesign  the  blades  for  the  conditions 
given  in  Exercise  18  when  the  nozzle  loss  is  3  per  cent,  of  the  theoretical 
velocity  developed  and  the  blade  losses  are  obtained  from  Fig.  51,  page  85. 

Observe  and  discuss  the  change  in  blade  angles  caused  by  including  the 
losses  in  the  design. 

Calculate  (i)  the  work  done  in  foot-pounds  per  second  per  pound  of 
steam;  (2)  the  steam  consumption  per  horsepower-hour  and  the  efficiency 
of  the  turbine. 

If  the  speed  of  the  turbine  is  20,000  revolutions  per  minute,  find  the 
diameter  of  the  mean  blade  circle. 

If  five  nozzles  are  used  for  a  maximum  load  of  50  horsepower,  find  the 
diameter  at  the  throat  of  each  of  these  nozzles,  assuming  they  are  all  of  the 
same  size. 

Exercise  20.  Make  the  necessary  calculations  and  draw  velocity  diagrams 
and  neat  sketches  of  the  blades  for  an  impulse  turbine  having  two  pressure 
stages  and  two  rows  of  moving  blades,  that  is,  two  velocity  stages  in  each 
pressure  stage,  for  the  following  requirements: 

The  initial  pressure  of  the  steam  supplied  to  the  turbine  is  165  pounds 
per  square  inch  absolute  and  is  expanded  in  the  first  set  of  nozzles  to  20 
pounds  per  square  inch  absolute.  In  the  second  set  of  nozzles  the  pres- 
sure falls  from  20  pounds  per  square  inch  absolute  to  2  pounds  per  square 
inch  absolute  (about  26  inches  vacuum).  The  nozzle  angles  are  20  degrees 
and  the  peripheral  speed  of  the  blades  or  buckets  js  500  feet  per  second,  the 
nozzle  loss  is  2  per  cent,  of  the  theoretical  velocity,  and  the  blade  losses  are 
to  be  taken  from  Fig.  51.  Assume  that  the  windage,  leakage,  and  bearing 
losses  amount  to  30  per  cent,  of  the  energy  developed  by  the  action  of  the 
steam  in  the  blades. 


368  APPENDIX 

The  rating  of  the  turbine  is  to  be  for  100  horsepower  at  1800  r.p.m. 
Calculate  the  number  of  buckets  and  the  height  of  the  buckets  for  the 
first  row  in  the  first  stage  and  for  the  last  row  in  the  second  stage. 

Observe  that  the  height  of  the  blades  for  the  first  row  in  each  stage  is 
determined  by  the  height  of  the  nozzles  which  discharge  into  the  blades. 

Exercise  21.  Design  the  blading  of  a  reaction  turbine  for  the  same  con- 
ditions given  in  the  second  exercise  on  page  108,  except  that  the  initial 
steam  pressure  is  to  be  165  pounds  per  square  inch  absolute,  and  tjie  final 
pressure  i  pound  per  square  inch  absolute. 

Exercise  22.  Design  a  combined  impulse  and  reaction  turbine,  taking 
the  general  data  the  same  as  for  the  preceding  exercise  and  the  expansion 
in  the  impulse  section  to  be  from  165  pounds  per  square  inch  absolute  to 
40  pounds  per  square  inch  absolute.  The  expansion  in  the  reaction  blading 
is  to  be  from  40  pounds  per  square  inch  absolute  to  i  pound  per  square  inch 
absolute. 

Sketch  the  blades  for  the  impulse  section,  assuming  there  are  two  velocity 
stages,  and  also  the  blades  for  the  first  and  last  stages  in  the  "reaction 
section." 

Exercise  23.  Determine  the  velocity  loss  in  feet  per  second  in  a  nozzle 
having  98  per  cent,  efficiency  at  its  proper  expansion,  which  is  from  125 
pounds  per  square  inch  absolute  pressure  to  28  inches  vacuum  (referred  to 
30  inches  barometer)  when  used  for 

(1)  165  pounds  per  square  inch  absolute  and  29  inches  vacuum. 

(2)  9  pounds  per  square  inch  absolute  and  26  inches  vacuum. 

State  also  the  corresponding  energy  loss  in  B.T.U.  per  pound  of  steam 
in  each  case,  and  by  what  percentage  the  efficiency  of  the  Rankine  cycle 
will  be  affected.  By  what  percentage  would  the  steam  consumption  of  a 
commercial  type  of  turbine  be  affected?  In  all  cases  mentioned  the  steam 
is  initially  dry  saturated. 


INDEX 

PAGE 

Absolute  velocity 68 

Accumulator,  Rateau 251-254 

Adiabatic  expansion * 21-24 

Adjustment  bearing 156,  175 

Allgemeine    Electrizitats    Gesellschaft    Turbines    (Curtis   and    Riedler- 

Stumpf) 182,  207 

Allis-Chalmers-Parsons  turbines 134,  167-1 70 

blades 161-163 

governor 234,  246 

rotor 320 

shroud  rings i6n 

Alternators,  parallel  operation 299 

with  forced  ventilation 361 

Area  of  floor  for  engines  and  turbines 299-304 

nozzles 30-41,  53 

Auxiliary  machinery,  power  for 296 

Available  energy 21-26,  54 

Balance  pistons 156,  167, 175 

Bearing  friction 92, 125,  150,  152 

Blades,  clearances  of 109-1 1 1 

conditions  of  best  efficiency 69-80 

design  of 81-113 

erosion  of 114 

impulse  and  reaction  upon 58-62 

radial  leakage  of 106-108 

materials  for 111-114 

c        peripheral  speeds  of 97,  100 

rotation  loss  of 117 

Bleeder  steam  turbines 2586 

Branca's  turbine 5 

British- Westinghouse  turbine 167 

British  thermal  unit 12 

Brown-Boveri-Parsons  turbine 167,  231,  241,  245 

Buckets  (see  also  Blades) 182,  207,  217 

By-pass  governing 240-245 

Centrifugal  force 313 

Chart,  entropy  —  total  heat  (in  Appendix). 

Clearances  of  blades,  axial no 

radial 109 

Commercial  testing  of  turbines 263-266 

Comparison  of  turbines  and  engines 1 26-13 1>  J37 

369 


370  INDEX 

PACK 

Composition  of  blades  ............  4  ...............................     111-114 

Condensing  water,  quantity  of.  ........................................      280 

Condensers  and  auxiliaries,  cost  of  .................................     293-296 

power  for  .......................................................     297 

Corrections  for  economy  curves  ........  .  .......................     126-136,  287 

for  rotation  losses  ...........................................  ....     118 

for  steam  turbine  tests  ........................................     1  26-136 

Cost  of  engines,  turbines,  and  auxiliaries  ...........................     291-296 

maintenance  and  operation  ...................................     290-291 

Critical  speed  of  rotating  shafts  .................  .  ...........     8,  147,  338-339 

Curtis  turbines  ..................................................     182-194 

analysis  of  losses  ................................................     194 

blades  ..................................................     113,  182,  185 

•diaphragms  .....................................................     182 

emergency  stop  valve  .........................................     190-191 

governor  ........................................     191,  226-229,  243-245 

manufacturers  ...................................................     182 

nozzles  ......................................................     27,  184 

oiling  system  ................................................     3°6~3  1  1 

small  sizes  ......................................................     192 

speeds  ..........................................................     19° 

steam  consumption  ...................................     133-136,  193,  266 

step  bearing  .....................................................     187 

superheat  corrections  .........................................     133,  193 

tests  on  .................................................     133-136,  266 

vacuum  corrections  .................................     132-134,  193,  278 

valve  gear,  electric  type  ...........................................     226 

hydraulic  ...................................................     229 

mechanically  operated  ....................................     226-229 

wheels  ......................................................     184-186 

Dake  turbine  ........................................................     217 

De  Laval  turbine  ............................................     5~6>  I40-i52 

analysis  of  losses  .................................................     I52 

bearings  ........................................................      J47 

blades  ......................................................     "3.  H6 

disks  ...........................................................     142 

gears  ...........................................................      J4& 

governor  ................................................      *5°»  220>  245 

nozzles  ....................................................     3>  27>  J44 

shaft  ...........................................................     J4» 

speeds  of  .......................................................      J44 

steam  consumption  ...............................................     J52 

superheat  corrections  ........................  .  ....................     I5° 

tests  on  .........................................................     27° 


vacuum  corrections 


II 


wheels  ..........................................................     *42 

Depreciation  of  power  plants  ......................................     294-295 


INDEX  371 

PACK 

Design  of  power  stations 297-306 

Design  of  turbines,  examples  in 86-ioS 

entropy-heat,  diagram  applied  to 87 

of  blades 81-83 

of  steam  nozzles 29~57 

Deterioration  of  turbines  and  auxiliaries 294-295 

Diaphragms Sr,  182,  196 

Disk  rotation  losses 117 

stresses  in 322-338 

Drums,  stresses  in 315-320 

Economy  of  engines,  best 288 

compared  with  turbines 285-288 

with  varying  superheat,  vacuum,  and  pressure 289 

Economy  of  small  engines  and  turbines 289-290 

Economy  of  standard  turbines 287 

Efficiency  of  blades 85 

steam  engines,  mechanical ^     286 

steam  nozzles 49,  86 

thermal  unit  basis  of 270-271 

turbines 85-86,  270 

turbine-generators 362 

Electric  generators  for  steam  turbines 358-362 

Entropy 17 

of  saturated  steam 18-23 

of  superheated  steam 18,  54-57 

of  water 19-20 

temperature  diagrams J7~25 

total  heat  chart 34,  37,  Appendix 

P>osion  of  blades 114 

Experiments  with  nozzles 29~33>  42~53 

on  flow  of  superheated  steam 52~53 

with  governors 245-246 

Flash-over  limit  of  turbine-generators 360 

Floor  area  for  engines  and  turbines 299-304 

Flow  of  steam: 

experiments  on 29,  52 

Grashof's  law  for 29~32 

Napier's  formula  for 29~3O 

saturated 29~33 

superheated 52~53 

weight  of  steam  flowing 29~33 

Fullagar's  blading 161-163 

Gas  turbines 34°~357 

compared  with  gas  engines 349 

Gauging  of  Parsons  blades 101 

Generators,  efficiency  of 362 

Glands,  water  packed 158,  209 


37 2  INDEX 

PAGB 

Governing  of  turbines,  methods  of 218-246 

by-pass ' 240-245 

cutting  out  nozzles 221-229 

experimental  data 245-246 

throttling 218-221 

varying  time  of  admission 229-240 

Governors,  turbine: 

Allis-Chalmers 234 

Brown-Boveri 231-232,  241-243 

Curtis „ 226-229,  243 

De  Laval 220-221 

Westinghouse-Parsons 233-234 

Wilkinson 238-240 

Grashof 's  law  for  flow  of  steam 20-32 

Hamilton-Holzwarth  turbine 204,  205 

Heat  diagram,  entropy Appendix 

applied  to  design 87 

Heat  theory 10-28 

Heat  units. ...- 12 

Heat,  mechanical  equivalent  of 13 

specific 12,  13,  54-56 

Hero's  turbine. 4,  8 

Horse-power,  internal  or  indicated 271 

conversion  table  for  kilowatts  to  brake 286 

Impulse  turbines 62-67 

and  reaction,  combined  (Westinghouse) 1 73-181 

blade  design 81-95 

blade  efficiency 69-77 

blade  losses 84-85 

comparison  with  reaction  turbines 108 

distinguished  from  reaction.. ....................................     58—62 

shape  of  blades 62 

Initial  velocity  of  steam 16 

Injection  water  (see  Condensing  water). 

Intermediate  (stationary)  blades 64,  182 

Jets,  impulse  and  reaction  of 58-60 

Kerr  turbine 211-214 

governor 213 

nozzles 212 

Kilowatts,  conversion  of,  to  brake  horse-power 286 

Knoblauch  and  Jakob's  specific  heat  of  superheated  steam 54~j6 

Labyrinth  packing 156,  167,  204,  209 

Lasche's  method  for  rotation  losses 121-1 24 

Leakage  of  steam  through  blades 125 

through  diaphragms 85,  115,  125,  204 

Losses  in  a  turbine 85,  92,  115-125,  152 

in  a  De  Laval  turbine,  analysis 152 


INDEX  373 

PAGE 

Low-pressure  turbines 247,  258 

accumulators  for 251-254 

combined  with  gas  engines 258,  291 

steam  consumption  of 251,  254-257 

Lubrication  of  turbines 306-3 1 1 

cost  of  oil  required 291 

of  Curtis  turbines 311 

Maintenance  and  operation 290-291 

Marine  turbines 249-250,  259-260 

designing  constants 97 

working  with  reciprocating  engines 249 

Materials  and  permissible  stresses  in  turbine  disks 337 

Mechanical  equivalent  of  heat 13 

losses  in  turbines 1 15-125 

Mixed  pressure  turbines 258a 

Moisture  correction  for  rotation  losses n8 

Motion,  absolute  and  relative 68-69 

Multicellular  turbines 195-202 

Napier's  formula  for  flow  of  steam 29-30 

Non-expanding  nozzles 47)  50-51,  173,  196 

Nozzles: 

area  of 30-41,  53 

design  of. 29—57,  94 

efficiency 49-50 

examples  of 27 

expansion  ratio 36-41 

expansion  wedges  in 217 

flow  of  steam^through 29~33,  42-45,  52 

length  of 46-47 

non-expanding 47,  50-51,  173,  196 

types  of 3)2 7-28 

losses  in 49-50 

Oil  pump,  Westinghouse  rotary 309 

Oiling  systems 306-3  r  i 

Curtis 3H 

Over-  and  under-expansion  in  nozzles 49-50 

Packing  glands 158 

Parsons  turbines 6-9,  153-172 

bearings 8,  164 

blades  and  lashing 158 

design  of 95-106 

governors  and  valve  gears 233-234,  245 

history 6-9 

manufacturers  of 167 

number  of  stages 97,  165 

pressure  diagram  (indicator) 235 

shroud  rings 161 

speeds  of 97 


374  INDEX 

Parsons  turbines  —  Continued  PACK 

steam  consumption  of 133-136,  170-171,  266-269 

superheat  corrections 133,  172 

vacuum  corrections 133-134,  172,  278 

Pelton  types  of  turbines 207-2 1 7 

Performance,  thermal  unit  basis  of 270-271 

of  engines  and  turbines 285-288 

Piping  of  turbine  stations 305-306 

for  superheated  steam 311-312 

Power  plant  economics 290-291 

Pressure  (initial  steam)  effect  on  economy 284 

corrections  for  economy  curves 126-136 

in  throat  of  nozzles 29~33 

stages 65 

Pressure-volume  diagrams 1 5-1 7 

Prices  of  turbines  and  auxiliaries 291-294 

Quality  of  steam '. 34~39 

Rankine  cycle  efficiency 28,  93,  272,  364,  368 

Rateau  accumulator. . 251-254 

Rateau  multicellular  turbines 195-202 

diaphragms 196 

governor  tests 245-246 

low  pressure 202 

manufacturers 202 

tests  on 255-269 

Reaction  of  jets 58-60 

Reaction  turbine 67 

blade  design 95-106 

blade  efficiency 77-80 

blade  losses 84-85 

distinguished  from  impulse 62 

Relative  velocity 68-69 

Riedler-Stumpf  turbines 207 

Rings,  stresses  in 3 14-316 

Rosenhain's  tests  on  flow  of  steam 47 

Rotation  losses 115-125 

factors  to  correct  for  superheat  and  moisture 118 

Rotors,  stresses  in 315-320 

Screw  turbine - o/ 

Shroud  rings 161,  185 

stresses  in 314-316 

Space  occupied  by  engines  and  turbines 299-304 

Sparking  limit  of  generators 358 

Specific  heat 12 

of  superheated  steam 13 

Specific  volume  of  superheated  .steam 53>  99 

Speed  of  turbines,  effect  on  economy  and  output 138,  284-285 

Stages  of  turbines, 64—65 


INDEX  375 

PACK 

Stages  of  turbines  —  Continued 

leakage  between 85,  115,  125,  204 

Steam  turbine  economics 273-312 

Steam  turbines: 

compared  with  steam  engines.. i,  13-17 

cost  of 291-294 

Step-bearing  of  Curtis  turbines 187 

Stodola's  experiments  with  nozzles 48-50 

Stresses  in  rotating  rings,  drums,  and  disks 3X3~338 

at  right  angles 320-322 

Sturtevant  turbine 207-21 1 

bearings 210 

buckets 207 

nozzles 208 

velocity  stages „ 210 

wheels. 211 

Superheated  steam  corrections  for  economy  curves 126-136,  287 

effect  on  economy 281-284 

flow  through  nozzles 52~53 

specific  heat  of 55~56 

specific  volume  of 53,  99,  118 

total  heat  of 271 

Taper  of  De  Laval  nozzles 46 

Temperature-entropy  diagrams J7~57 

Temperature 10— n 

absolute 1 1 

Terry  turbine 214-216 

Tests  on  turbines 132-139,  261-271 

Thermal  units,  definition  of 12 

basis  of  performance 270-271 

Thermodynamic  efficiency 271-272 

Thermometer  correction n 

Thrust,  end,  in  Parsons  turbine 156 

Torque  line 138 

Total  heat  of  superheated  steam 271 

Types  of  steam  turbines 58-62 

Vacuum,  condensing  water  required  for 275,  280 

corrections  for  economy  curves 1 26-136,  278,  280 

effect  on  engine  and  turbine  economy 273-281,  288 

most  profitable „ 273 

Vanes  (see  Blades). 

Velocity,  absolute 68-69 

diagrams 63,  66-70 

relative 68-69 

Velocity  of  steam: 

calculation  of 24,  26 

stages 64,  74, 1 73,  210 


376  INDEX 

Volume,  specific:  PAGE 

at  high  vacuums  (table) 277 

of  superheated  steam 53,  99 

Water-packed  glands 157,  158,  209 

Water-rate  curves 262 

Westinghouse  oil  pump 309 

Westinghouse  impulse  and  reaction  turbines 1 73-181 

adjustment  bearing 175 

advantages  of 181 

blades  of  impulse  section 173 

emergency  speed  limit 177-180 

nozzles  and  nozzle  block 177-178 

reduction  gear 26oa 

velocity  stages 173 

Westinghouse-Parsons  turbines 165-166 

auxiliary  admission  valve  (overload) 234 

bearings 164 

blades  and  lashing 159 

governor  and  valve  gear 233-234,  245 

indicator  diagram 235 

tests  on 133-136,  171,  255,  267-269,  271 

water-packed  glands 157-158 

Wilkinson  turbines 202,  204 

governor  and  valve  gear 202 

stage  packing 204 

Willans  and  Robinson  turbine 161 

Willans  lines 99,  1 24 

Zoelly  turbines 205,  206 

blades  of 206 

governor 245-246 

nozzles 205 

tests  on ; . . .     270 

wheels. . .  206 


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